EXPERIMENTAL INVESTIGATIONS OF NH3/CO2 CASCADE SYSTEM … · variations of the system solution are...

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EXPERIMENTAL INVESTIGATIONS OF NH3/CO2 CASCADE SYSTEM FOR SUPERMARKET REFRIGERATION

by

Arash Soleimani Karimabad

Master of Science Thesis Master Program of Sustainable Energy Engineering

2006

Department of Energy Technology Royal Institute of Technology

Stockholm, Sweden

Master Thesis / Arash Soleimani Karimabad

ABSTRACT A refrigeration system solution for a supermarket has been built in IUC laboratory in Katrineholm. The system is equipped with extensive instrumentations to collect data and perform online diagnosis. Several variations of the system solution are applied for validation and possible modifications. In the project period we investigated and evaluated the performance of NH3/CO2 cascade system for supermarket refrigeration application in a laboratory controlled environment. The system solution under investigation replicates a medium size supermarket in Sweden. CO2 is used at the low and medium temperature levels with several possibilities available for system variations and parametric analysis. Overall system evaluation has been performed. Different circulation rates at the medium temperature level have been tested. Control system of the freezing cabinets has been evaluated. The CO2 compressor and cascade condenser performance have also been investigated. The possibilities for gravity circulation have been examined. The results of the tests show that CO2 system has a reasonable COP with good agreement between the experimental data, computer model, and similar systems in real installations. Evaluating the different circulation rates of CO2 at the medium temperature level shows that it should be kept as low as possible and there is no optimum value as in conventional brine in indirect systems. The control of the electronic expansion valve should be based on the evaporating pressure rather than measuring surface temperature of the evaporator. Since a high instability observed in keeping low superheat value in the freezer cabinets, using an internal heat exchanger was avoided not to have high discharge temperature. The CO2 compressor has reasonable performances with good isentropic and volumetric efficiencies. The heat transfer characteristics in the cascade condenser are almost the same while varying between forced condensation and thermo-syphon scenarios. In the gravity circulation test, mass flow of CO2 pulses and temperatures fluctuates accordingly which indicates that the available height is not sufficient enough to maintain continuous flow of CO2. The same trend is observed at all simulator capacities. The pressure drop in the pipes and the heat exchangers are quite small and the main pressure drop sources in the CO2 system are the filters and the mass flow meter which is not overcome by the limited available head. Based on the experience in designing the system, running and evaluating its modifications, it is possible to conclude the application of a NH3/CO2 cascade refrigeration system would be a comparable environmental friendly and efficient option for a medium size supermarket in Sweden. It will be possible to point out the components in the system that have low performances and


that will direct the research in the future to improve the efficiency of such elements in the systems. This thesis is based on the material of the final report of the first phase of the CO2 project at IUC Katrineholm in cooperation with KTH [1]. Authors of the report are: Sawalha, S., Soleimani, A., Rogstam, J.


ACKNOWLEDGMENTS The project “CO2 in Supermarket Refrigeration” is a group effort among IUC, KTH/Applied Thermodynamic and Refrigeration division, Ahlsell, Huurre, AGA, WICA and ICA. The project is financed by Energimyndigheten (STEM) and the involved companies. I would like to extend my appreciation for the positive spirit of cooperation from all the project partners. Many people have contributed to carry out this work and to whom I wish to express my gratitude. First of all, I would like to thank my supervisor Professor Björn Palm for his guidance and valuable advices and especially I am highly grateful to Jörgen Rogstam and Samer Sawalha for their great technical and moral support, their useful comments, and their friendship. Special thanks go to Lars-Åke Johansson for his great help, many valuable comments, and stimulating discussions. I would like to extend my gratitude to Per-Olof Nilsson because of his great help in IUC refrigeration laboratory and also all the people at IUC, Göran Lundin, Ann-Britt Rasmussen, Katarina Mjörnebrant. I would also thank you Nina Schröder from AGA Gas AB and Peter Nordén from KTH energy department for their great help in facilitation in my scholarship procedure. Finally I want to express my gratitude to my beloved wife for all her mental support and understanding. Arash Soleimani March 2006


TABLE OF CONTENTS ABSTRACT ............................................................................................................................................2 ACKNOWLEDGMENTS......................................................................................................................5 TABLE OF CONTENTS .......................................................................................................................6 LIST OF FIGURES................................................................................................................................7 NOMENCLATURE AND DEFINITION.............................................................................................9 1 INTRODUCTION.......................................................................................................................10

1.1 SUPERMARKET REFRIGERATION ...........................................................................................11 1.2 NATIONAL LAWS, REGULATIONS, AND STANDARDS ............................................................11 1.3 NATURAL REFRIGERANTS ....................................................................................................12 1.4 CARBON DIOXIDE AS A REFRIGERANT .................................................................................13

1.4.1 Different System Configurations for CO2 in Supermarket Applications ........................14 1.4.1.1 Cascade System .................................................................................................................. 14 1.4.1.2 Secondary Refrigerant Application..................................................................................... 15 1.4.1.3 Transcritical Cycle .............................................................................................................. 15

1.4.2 Advantages and Disadvantages for Carbon Dioxide......................................................16 2 OBJECTIVE ...............................................................................................................................17 3 CASCADE REFRIGERATION RIG IN IUC..........................................................................17

3.1 SYSTEM OVERVIEW .............................................................................................................17 3.1.1 System solution ...............................................................................................................18 3.1.2 System layout ..................................................................................................................20

4 MEASUREMENT FACILITIES...............................................................................................21 5 EXPERIMENTAL RESULT.....................................................................................................23

5.1 OVERALL SYSTEM ANALYSIS................................................................................................23 5.1.1 Load Measurements........................................................................................................23 5.1.2 Energy Balance Test .......................................................................................................24

5.1.2.1 Cascade condenser cooling capacity ................................................................................... 26 5.1.2.2 Ammonia compressor capacity and efficiencies ................................................................. 30 5.1.2.3 CO2 compressor capacity and efficiencies.......................................................................... 35

5.1.3 System Efficiency ............................................................................................................42 5.1.3.1 Low stage COP ................................................................................................................... 42 5.1.3.2 High stage COP................................................................................................................... 44 5.1.3.3 Total COP ........................................................................................................................... 45

5.2 SYSTEM VARIATIONS ...........................................................................................................48 5.2.1 Gravity Circulation and Pump Circulation Ratio...........................................................49 5.2.2 Cascade Condenser ........................................................................................................58 5.2.3 Flashing Gas in Liquid Lines .........................................................................................63 5.2.4 Freezing Cabinets Control .............................................................................................65

6 CONCLUSION AND DISCUSTION ........................................................................................73 7 REFERENCES............................................................................................................................75


LIST OF FIGURES FIGURE 1 – FRICK’S “SPLIT-STAGE” AMMONIA/CARBON DIOXIDE SYSTEM [4].........................................10 FIGURE 2 – COP OF CO2 COMPARED TO OTHER MAIN REFRIGERANTS IN LOW STAGE (A) AND IN HIGH

STAGE (B) OPERATIONS [6].............................................................................................................14 FIGURE 3 – TRANSCRITICAL CYCLE IN THE CO2 PRESSURE–ENTHALPY DIAGRAM ..................................15 FIGURE 4 – SCHEMATIC DIAGRAM OF THE CO2 CIRCUIT IN THE NH3/CO2 CASCADE SYSTEM WITH CO2

AT THE MEDIUM TEMPERATURE LEVEL ..........................................................................................18 FIGURE 5 – SCHEMATIC DIAGRAM OF NH3/CO2 CASCADE SYSTEM WITH CO2 AT THE MEDIUM

TEMPERATURE LEVEL ....................................................................................................................19 FIGURE 6 – DETAILED SCHEMATIC DIAGRAM OF THE NH3/CO2 CASCADE SYSTEM TEST RIG..................21 FIGURE 7– ERROR IN % OF ACTUAL MASS FLOW RATE WITH 95% CONFIDENCE (PROBABILITY)..............22 FIGURE 8 – SCHEMATIC OF THE ACTIVE LINES AND COMPONENTS IN THE ENERGY BALANCE TEST..........24 FIGURE 9 – PLOT OF THE TEMPERATURES AT THE SYSTEM BOUNDARIES DURING THE ENERGY BALANCE

TEST PERIOD [5] .............................................................................................................................25 FIGURE 10 – ENERGY BALANCE WITH FIXED VALUE OF VOLUMETRIC EFFICIENCY [5].............................26 FIGURE 11 – ENERGY BALANCE WITH CALCULATED VALUE OF VOLUMETRIC EFFICIENCY [5].................28 FIGURE 12 – VOLUMETRIC EFFICIENCY OF THE AMMONIA COMPRESSOR [5] ...........................................29 FIGURE 13 – ENERGY BALANCE WITH ADJUSTMENT TO THE CALCULATED VALUE OF VOLUMETRIC

EFFICIENCY [5]...............................................................................................................................30 FIGURE 14 – MEASURED AND CALCULATED ISENTROPIC EFFICIENCIES OF THE AMMONIA COMPRESSOR [5]

......................................................................................................................................................31FIGURE 15 – TYPICAL VALUES OF ELECTRIC MOTOR EFFICIENCY VERSUS THE MOTOR SHAFT POWER [6] 32FIGURE 16 – CALCULATED AMMONIA COMPRESSOR SHAFT POWER AND MEASURED ELECTRIC MOTOR

EFFICIENCY [5]...............................................................................................................................32 FIGURE 17 – CALCULATED ELECTRIC MOTOR EFFICIENCY FOR THE ENERGY BALANCE TEST [5] .............33 FIGURE 18 – CALCULATED ELECTRIC MOTOR EFFICIENCY FOR AMMONIA COMPRESSOR WITH ALL

CYLINDERS RUNNING [5]................................................................................................................34 FIGURE 19 – CALCULATED ELECTRIC MOTOR EFFICIENCY FOR AMMONIA COMPRESSOR AT DIFFERENT

MOTOR SPEEDS [5] .........................................................................................................................34 FIGURE 20 – CALCULATED OLD CO2 COMPRESSOR SHAFT POWER AND MEASURED ELECTRIC MOTOR

EFFICIENCY [5]...............................................................................................................................35 FIGURE 21 – CALCULATED AND MEASURED OLD CO2 COMPRESSOR DISCHARGE TEMPERATURE [5] ......36 FIGURE 22 – ISENTROPIC EFFICIENCY OF THE CO2 COMPRESSOR AT DIFFERENT PRESSURE RATIOS ........37 FIGURE 23 – MEASURED ISENTROPIC AND VOLUMETRIC EFFICIENCIES OF THE OLD CO2 COMPRESSOR [5]

......................................................................................................................................................38FIGURE 24 – SCHEMATIC DIAGRAM OF THE SYSTEM SHOWS THE LEAKING LINE [5] ................................39 FIGURE 25 – EVAPORATING AND CONDENSING TEMPERATURES FOR THE TEST IN NEW CO2 COMPRESSOR

......................................................................................................................................................39FIGURE 26 – POWER CONSUMPTION OF THE NEW CO2 COMPRESSOR AND THE COOLING CAPACITY ........40 FIGURE 27 – THE DISCHARGE TEMPERATURE OF THE NEW CO2 COMPRESSOR ........................................41 FIGURE 28 – ISENTROPIC AND VOLUMETRIC EFFICIENCIES OF THE CO2 COMPRESSOR WITH THE PRESSURE

RATIO PLOTTED..............................................................................................................................41 FIGURE 29 – CALCULATED AND MEASURED LOW STAGE COP FOR OLD CO2 COMPRESSOR [5]...............43 FIGURE 30 – CALCULATED AND MEASURED LOW STAGE COP FOR NEW CO2 COMPRESSOR ...................44 FIGURE 31 – CALCULATED AND MEASURED AMMONIA UNIT COP [5] .....................................................45 FIGURE 32 – CALCULATED AND MEASURED TOTAL COP WITH THE OLD CO2 COMPRESSOR [5] .............46 FIGURE 33 – MEDIUM AND LOW STAGE COOLING LOAD IN THE TEST WITH NEW CO2 COMPRESSOR......47 FIGURE 34 – EVAPORATING AND CONDENSING TEMPERATURE FOR BOTH LOW STAGE AND HIGH STAGE

IN THE TEST WITH NEW CO2 COMPRESSOR....................................................................................47 FIGURE 35 – MEASURED TOTAL COP, HIGH STAGE COP, LOW STAGE COP, WITH THE NEW CO2

COMPRESSOR AT HIGHER COOLING LOADS .....................................................................................48 FIGURE 36 – SCHEMATIC DIAGRAM OF THE CIRCULATION RATE TEST CIRCUIT........................................49 FIGURE 37 – TEMPERATURES AROUND THE LOAD SIMULATOR AND THE MASS FLOW OF CO2 DURING THE

GRAVITY CIRCULATION TEST ON THE LOAD SIMULATOR ................................................................50 FIGURE 38 – MEDIUM CABINET TEMPERATURES AND THE MASS FLOW OF CO2 DURING THE GRAVITY

CIRCULATION TEST ON THE MEDIUM CABINET ...............................................................................51 FIGURE 39 – SIMULATOR’S PRESSURE DROP AND CIRCULATION RATIO FOR SHORT WAITING TIME TEST [5]

......................................................................................................................................................52


FIGURE 40 – SIMULATOR’S PRESSURE DROP AND CIRCULATION RATIO FOR LONG WAITING TIME TEST [5]......................................................................................................................................................53

FIGURE 41 – SIMULATOR’S PRESSURE DROP AT DIFFERENT CIRCULATION RATIO FOR SHORT AND LONG WAITING TIME [5] ..........................................................................................................................54

FIGURE 42 – TEMPERATURES AROUND SIMULATOR [5] ...........................................................................55 FIGURE 43 – SIMULATOR’S LMTD AT DIFFERENT CR FOR LONG AND SHORT WAITING TIME TESTS [5] ..56FIGURE 44 – MEDIUM CABINET’S PRESSURE DROP AND LMTD VERSUS DIFFERENT CIRCULATION RATIO

FOR GRAVITY CIRCULATION TEST ..................................................................................................57 FIGURE 45 – TEMPERATURES AROUND THE MEDIUM CABINET ................................................................58 FIGURE 46 – SCHEMATIC DIAGRAM FOR INDIRECT OR THERMOSYPHON ARRANGEMENT: (MEDIUM RETURN

TO THE TANK) ................................................................................................................................59 FIGURE 47 – TEMPERATURES ACROSS CASCADE CONDENSER FOR INDIRECT OR THERMOSYPHON

ARRANGEMENT IN MEDIUM CIRCUIT ..............................................................................................59 FIGURE 48 – SCHEMATIC DIAGRAM FOR DIRECT OR FORCED CONDENSATION ARRANGEMENT: (MEDIUM

RETURN TO THE CASCADE CONDENSER).........................................................................................60 FIGURE 49 – TEMPERATURES ACROSS CASCADE CONDENSER FOR DIRECT OR FORCED CONDENSATION

ARRANGEMENT IN MEDIUM CIRCUIT ..............................................................................................61 FIGURE 50 – SCHEMATIC DIAGRAM FOR INDIRECT OR THERMOSYPHON ARRANGEMENT: (DISCHARGE HOT

GAS TO THE TANK) .........................................................................................................................61 FIGURE 51 – TEMPERATURES ACROSS CASCADE CONDENSER FOR INDIRECT OR THERMOSYPHON

ARRANGEMENT IN LOW TEMPERATURE CIRCUIT ............................................................................62 FIGURE 52 – SCHEMATIC DIAGRAM FOR DIRECT OR FORCED CONDENSATION ARRANGEMENT:

(DISCHARGE HOT GAS MIXING WITH THE SAT. VAP. FROM THE TANK THEN TO THE CASCADE CONDENSER) ..................................................................................................................................62

FIGURE 53 – TEMPERATURES ACROSS CASCADE CONDENSER FOR DIRECT OR FORCED CONDENSATION ARRANGEMENT IN LOW TEMPERATURE CIRCUIT ............................................................................63

FIGURE 56 – BASIC SCHEMATIC OF INTERNAL HEAT EXCHANGER SOLUTION...........................................64 FIGURE 57 – BASIC SCHEMATIC OF THE SOLUTION WHERE HEAD IS ADDED TO THE LIQUID BEFORE

EXPANSION VALVE.........................................................................................................................64 FIGURE 58 – BASIC SCHEMATIC OF THE FREEZER’S EVAPORATOR WITH THE MEASURING POINTS

INDICATED .....................................................................................................................................65 FIGURE 59 – TEMPERATURES AROUND THE FREEZER CABINET WITH TEMPERATURE BASED CONTROLLER.

SUPERHEAT SET VALUE IS 3ºC. [5].................................................................................................66 FIGURE 60 – TEMPERATURES AROUND THE FREEZER CABINET WITH TEMPERATURE BASED CONTROLLER.

SUPERHEAT SET VALUE IS 7ºC. [5].................................................................................................67 FIGURE 61 – BASIC SCHEMATIC OF THE FREEZER’S EVAPORATOR WITH THE MEASURING POINTS

INDICATED .....................................................................................................................................67 FIGURE 62 – TEMPERATURES AROUND THE FREEZERS AND SIMULATOR. SUPERHEAT SET VALUE IS 8ºC.

[5]..................................................................................................................................................68 FIGURE 63 – TEMPERATURE VALUES INPUT TO THE CONTROLLER IN THE FREEZERS. SUPERHEAT SET

VALUE IS 8ºC. [5]...........................................................................................................................69 FIGURE 64 – TEMPERATURE VALUES INPUT TO THE CONTROLLER IN THE SIMULATOR. SUPERHEAT SET

VALUE IS 8ºC. [5]...........................................................................................................................69 FIGURE 65 – TEMPERATURES AROUND THE FREEZER WITH PRESSURE BASED CONTROLLER. SUPERHEAT

SET VALUE IS 5ºC. [5] ....................................................................................................................70 FIGURE 66 – TEMPERATURES AROUND THE FREEZER WITH TEMPERATURE BASED CONTROLLER.

SUPERHEAT SET VALUE IS 5ºC. [5].................................................................................................71 FIGURE 67 – TEMPERATURE VALUES INPUT TO THE PRESSURE BASED CONTROLLER IN THE FREEZER.

SUPERHEAT SET VALUE IS 5ºC. [5].................................................................................................72 FIGURE 68 – TEMPERATURE VALUES INPUT TO THE TEMPERATURE BASED CONTROLLER IN THE FREEZER.

SUPERHEAT SET VALUE IS 5ºC. [5].................................................................................................72 FIGURE 69 – TEMPERATURES AROUND THE SIMULATOR WITH SUPERHEAT SET VALUE OF 5ºC. [5] .........73


NOMENCLATURE AND DEFINITION CFC: (Chloroflourocarbon) any of various halocarbon compounds consisting of carbon, hydrogen, chlorine, and fluorine HCFC: (Hydrochlorofluorocarbons) are halogenated compounds containing carbon, hydrogen, chlorine and fluorine. They have shorter atmospheric lifetimes than CFCs and deliver less reactive chlorine to the stratosphere where the "ozone layer" is found CEN: Committee for European Normalization. A standards setting body including the members of the European Community COP: Coefficient of Performance

) ( ) ( int

) (

meterflowmassflowMassqmmeterflowmasserrorpoZeroZ

meterflowmassErrorE

===

m& : Refrigerant mass flow rate sη : Volumetric efficiency of the compressor

inρ : Density of the refrigerant at the inlet to the compressor

sV& : Swept volume flow

srV& : Compressor displacement volume

rn : Compressor rated speed n : Compressor speed Q& : Cooling capacity dh : Enthalpy difference across the heat exchanger dQ : Difference between total cooling capacity and the provided load dQaverage: average value for dQ

1P : Discharge pressure

2P : Suction Pressure

kη : Isentropic efficiency

1T : Condensing temperature

2T : Evaporating temperature

ShaftE& : Compressor shaft power

thermalη : Compressor thermal efficiency

elη : Compressor electrical efficiency

elE& : Compressor electrical power

compdh : Compressor enthalpy difference RPM: Round per Minute CR: Circulation Ratio IHX: Internal Heat Exchanger LMTD: Logarithmic Mean Temperature Difference


1 INTRODUCTION

The history of refrigeration backs to thousands of years ago in ice houses and keeping the ice for summer use. First application of ice boxes dates back to 19th century. Using the latent heat of evaporation of a fluid called refrigerant is the basic concept in refrigeration. The technology is associated with the pumping heat from a body or a fluid of a low temperature to the one with higher temperature. [2] In 1856, James Harrison introduced the first practical refrigerator in a closed vapor compression refrigeration cycle for keeping beer cold. Later on the first electric refrigerator manufactured in 1913 in US and the first widespread refrigerators were manufactured by General Electric in 1927. [2] Earliest refrigerator units were using the toxic refrigerant like ammonia (R717), sulfur dioxide (R764), and methyl chloride (R40). Carbon dioxide (R744) has also been used in vapor compression refrigeration cycle in closed refrigeration cycle, but it required very high pressure. It gained favor in marine application in 1880s as it was much more efficient than open air cycles which were widely used until then. Also it had the benefit from its non-toxic and non-flammable nature. Improving the technology for small and fast running compressor for ammonia as the opponent of carbon dioxide made the ammonia much more feasible than carbon dioxide system around 1920s. Still because of ammonia’s toxicity and pungent smell there was not a great tendency of its application. In response, Frick Company pioneered an installation of a hybrid system in 19th January 1932 which they called “split-stage”. [4]

Figure 1 – Frick’s “split-stage” ammonia/carbon dioxide system [4]

The system featured the opportunity of minimizing the charge of ammonia as the refrigerant in high stage and also the benefit of condensing carbon dioxide in a moderate pressure.


Soon after, the introduction of Dichlorodifluoromethane (R12) which happened in 1931 rapidly took the place of carbon dioxide and ammonia. They enjoyed their high efficiency and flexibility features comparing to ammonia and carbon dioxide. Between 1950 and 1970, introduction of new synthetic extensively substituted with the pre-introduced refrigerants but exploring the damaging effect of CFCs to ozone layer made 43 nations to sign Montreal Protocol in 1987 and reduce the production of CFCs afterward. Since then, the researches have been redirected back again to the application of old refrigerants. Advancements in manufacturing technology have aroused the hope to overcome obstacles in modern refrigeration techniques. [4]

1.1 Supermarket refrigeration

A supermarket can be defined as a self service grocery store in a larger scale with wider variety of product selections by the customers. Working in grocery stores were much more labor-intensive thus more expensive than a supermarket. So it gained great favor rapidly starting in early 20th century until now. Frozen foods like meat, poultry, and fish as well as chilled foods like fruits, and vegetables are supposed to be kept in presence of costumers in the cabinets in different certain cold temperatures. [3] Considering the basic refrigeration cycle, in supermarket applications, the difference between evaporating and condensing temperatures is quite large. Systems are usually operating at a constant condensing temperature of about 40ºC and the evaporating temperature of about -30ºC for the frozen products and -10ºC for the chilled products. Therefore, cascade, two or multi-stage system solutions become favorable and they are well adaptable for the two-temperature level requirement for chilled and frozen products in the supermarket.

1.2 National Laws, Regulations, and Standards

In January 1995, the Swedish parliament started banning new installation of CFC refrigerants. These environmental legislations have been provided to phase out CFC and HCFC refrigerants. The use of such refrigerants has been prohibited from January 2000. This has led to the investigation for finding new refrigeration system designs applying new environmental friendly refrigerants. [7] The supermarket sector has been affected by the substitution of CFC and HCFC refrigerants and by a key reconsideration of the energy utilization. To minimize the refrigerant leakage, indirect systems have been used as new solution in supermarkets in Sweden. Roughly 3% of total electricity is consumed in supermarkets which have been about 1.8 TWh/year in 2001. [7] Until now, there is no specific legislation on the application of CO2 in supermarket refrigeration. Generally it is well enough covered within the existing regulations for other refrigerants and more general health and safety regulations.[17, 18] For example regarding the practical limit calculations for


the build up of CO2 in an occupied space it is treated in the same way as R22 would be. A common issue for CO2 systems in supermarkets is the high pressure at standstill; it is always an important factor to be taken into account when considering CO2 as the working fluid. If the plant would be stopped for maintenance, component failure, power cut or any other reason, then the refrigerant inside the plant will start to gain heat from the ambient and the pressure inside the plant will consequently increase. One element of the safety standard EN378 that was inadequate was the determination of allowable pressure for the low temperature circuit in a cascade system. The early mentioned value was 12.4 bar a for low pressure side. This has been addressed by the CEN safety standards committee in the revised version of EN378 which is due to be published later this year. The new limit has not been published yet but it would certainly consider the new growing application of R744 as an alternative refrigerant. At this time it is not possible to quote the new standard, but it is possible to refer to the fact that it is under review. The other standard to refer to is the UK’s Institute of Refrigeration (IOR) Safety Code for Systems in utilizing carbon dioxide. This is available on ww.ior.org.uk. It is not a legal requirement in the UK, but it is a supporting document.

1.3 Natural Refrigerants

Natural working fluids are substances naturally existing in biosphere. They mostly have zero or negligible global environmental impacts. So they are considered as promising alternatives for CFC refrigerants. There are five substances that we can call “natural refrigerants”:

• Air • Water • Ammonia • Hydrocarbons • Carbon dioxide

Air has been used in a variety of gas cycles, with no change of phase, and can achieve reasonably low temperatures, but the low theoretical efficiency of the Brayton cycle and the difficulty of getting close to that ideal have limited its use. Water vapor has been used with large centrifugal and axial turbines in open systems but the low pressures, large swept volumes, and evaporation temperature limit of 0ºC place severe restrictions on its use and make it fundamentally unsuited to smaller air conditioning systems and industrial cooling and freezing applications. Ammonia, hydrocarbons, and carbon dioxide have a broader range of application, and are used in much more conventional systems. Despite a generally excellent safety record there is a strict limit on the allowable charge of hydrocarbon systems, which makes them unsuitable for use in large water


chillers and industrial systems unless relevant safety standards can be applied. In many ways ammonia is ideal for large industrial systems where its mild flammability, pungent smell and low threshold limit value do not present problems. It is, however, clearly unsuited to domestic, automotive and small commercial refrigeration and heat pump systems. This leaves carbon dioxide as the only natural refrigerant to find favor across the broad spectrum of automotive, domestic, commercial and industrial refrigeration and air-conditioning systems.

1.4 Carbon Dioxide as a Refrigerant

In the late 20th century, the CFC in refrigeration system started to be banned from further usage and the research in new alternative refrigerant has come to the interest. The researchers have been searching for old techniques such as the application of carbon dioxide as a coolant to substitute in refrigeration system. Among the alternatives for the new refrigerants, CO2 is rather non-toxic, non-flammable, non-explosive, and it is compatible to normal lubricants and common construction materials. Extensive research and development on CO2 technology has been performed internationally over the past decade since the revival of carbon dioxide. Most of the researches have been focusing on the usage of carbon dioxide in commercial refrigeration. Specifically a group in Europe has been founded in July 2000 with the name of Carbon Dioxide Interest Group (c-dig) initially to share the knowledge for industrial applications. Also IIR-Gustav Lorentzen series of conferences for natural refrigerants show the growing research interests in the field. Currently there are number of examples for different applications of carbon dioxide in refrigeration. For instance, CO2 has been used in heat pumps in residential heating, automotive heating, and water heating. In 1990 Prof Gustav Lorentzen published a patent application for a trans-critical carbon dioxide system for automotive air-conditioning. [16] There have afterward been lots of researches in this area and currently the automotive industries are increasingly working in usage of carbon dioxide in trans-critical cycle. In commercial refrigeration such as in shops and supermarkets, the main technology that was commercially utilized, was to use CO2 as secondary refrigerant in indirect systems for freezing temperature applications. The CO2 indirect circuit is connected to the primary refrigerant cycle via its evaporator, which evaporates the primary refrigerant on one side and condenses CO2 at the other side. Liquid circulation of CO2 in the indirect system is usually achieved via pump. Another technology in commercial refrigeration is the cascade configuration. Cascade systems with CO2 in the low temperature stage have also been applied in several supermarket installations and are becoming a more and more competitive alternative. Looking to thermo-physical properties of carbon dioxide, the low critical point for CO2 of 31ºC implies that it will operate with


better theoretical COP between low temperature ranges further below the critical point compared to high stage conditions. Therefore, the application of CO2 in the low stage of a cascade system yields a reasonable theoretical COP compared to other refrigerants, see Figures 2a and 2b. In Figure 2b, the COP values above the critical point are obtained at optimum pressure on the high pressure side.

-15 -10 -5 0 5 10 153

4

5

6

7

8

9

10

11

12

CO

P

CO2R404A

R134aAmmonia

Evaporating Temperature=-35 °C

Condensing temperature (°C)

15 20 25 30 35 40 451

2

3

4

5

6

7

8

9

10

Condensing temperature (°C)

CO

P CO2R404AR134aAmmonia

Evaporating Temperature=-10 °C

(a) (b) Figure 2 – COP of CO2 compared to other main refrigerants in low stage (a) and in high stage (b) operations [6]

Several factors contribute to improve the COP of CO2. The favorable thermo- physical properties of CO2 result in low pressure and temperature drops in the system. From the heat transfer point of view, the low surface tension will make boiling easier and therefore will improve the heat transfer. Also, due to the low pressure drop [11] the components of the system will be smaller while the mass flow rate of the refrigerant will be comparable to R404A, R22, R502 and R134a refrigerants, which will result in high mass flux of CO2 in the heat exchangers. Another improvement to the COP comes from the improved volumetric efficiency of the CO2 compressor compared to conventional refrigerants; this is due to the lower pressure ratio across the CO2 compressor. [6]

1.4.1 Different System Configurations for CO2 in Supermarket Applications

1.4.1.1 Cascade System

As briefly explained before, cascade system provides an interesting solution where CO2 can be used at the low temperature stage and another refrigerant can be used at the high temperature stage. In this case each refrigerant operates in the system’s boundaries where it can provide good COP. CO2 in such solution operates sub-critically between -37ºC evaporating and -8ºC condensing temperatures with acceptable pressure levels (11 and 28 bars


respectively). Conventional refrigeration components can be used in this case. Gaining experience and confidence in operating CO2 in indirect system solution made it possible to apply the same concept on the medium temperature level; this can be used in cascade or trans-critical solution. In this case the condensing CO2 from the lower stage at a temperature of -8ºC is accumulated in a tank and the CO2 in the tank is circulated in the medium temperature cabinets to provide the required cooling.

1.4.1.2 Secondary Refrigerant Application

CO2 has been successfully used in indirect system solution, the pumping power needed is quite small compared to conventional brine systems, and this is due to the small CO2 volume flow rate and the resulting low pressure drop. The small volume flow rate is due to the phase changing process on the CO2 side which also contributes to improving heat transfer on the refrigerant side compared to the cases with non-phase changing fluids, such as conventional brines.

1.4.1.3 Transcritical Cycle

Compared to conventional refrigerants, the most distinguishing property of CO2 is the low critical temperature of 31.1ºC. Vapor compression systems with CO2 operating at normal refrigeration, heat pump and air-conditioning temperatures will therefore work close to and above the critical pressure of 7.38 MPa. Heat rejection will, in most cases, take place at supercritical pressure, causing the pressure levels in the system to be high, and the cycle to be ‘transcritical’, i.e. with subcritical low-side and supercritical high-side pressure.

Figure 3 – Transcritical cycle in the CO2 pressure–enthalpy diagram

During operation at high ambient air temperatures the CO2 system will operate in a transcritical cycle most of the time heat rejection then takes place by cooling the compressed fluid at supercritical high-side pressure. The low-


side conditions remain subcritical, however, as shown in Figure 3, at supercritical pressure, no saturation condition exists and the pressure is independent of the temperature. In conventional subcritical cycles, the specific enthalpy in point 3 is mainly a function of temperature, but at supercritical high-side conditions the pressure also has a marked influence on enthalpy.

1.4.2 Advantages and Disadvantages for Carbon Dioxide

Among natural refrigerants and new alternatives, CO2 has a great safety features since it is non-flammable, non-toxic, and non-explosive. It has an excellent availability thus it is cheap while it doesn’t participate in ozone depletion. However increasing amount of CO2 as a major product for fossil fuel burning in the globe is supposed to be a main participant to global warming, its application in refrigeration would not contribute to global warming. It has negligible GWP and even zero if it is drawn from the waste product of industrial process. Also as explained before, CO2 is mostly compatible with a wide range of the construction materials and oils being used in refrigeration technology. In addition, competitiveness with respect to other refrigeration systems has been reached and proved in several applications through several experimental researches and real installations. Heat transfer to carbon dioxide is often characterized as being superior to other refrigerants. High vapor density, low surface tension by one order of magnitude, and low vapor viscosity considerably influence the convective and nucleate boiling characteristics of CO2. This results in heat transfer coefficient of CO2 that are greater than those of conventional refrigerant by 2-3 times at the same saturation temperature while its two phase pressure drops are significantly smaller. The low pressure drop and the low volume flow rate, due to the high CO2 vapor density, will also contribute to minimizing the energy consumption of the pump in the indirect circuit which will give CO2 a major advantage compared to brine based systems. A further advantage related to the use of CO2 is its higher volumetric capacity due to its high working pressures enabling small equipment components and small-diameter lines to be used. This has favored carbon dioxide in MAC application where there is a limited space available. In the transcritical cycle, gas cooler pressure and temperature are not linked as in the subcritical two-phase region. Taking advantage of rather high discharge temperature has favored carbon dioxide as an interesting refrigerant for a heat pump application such as in domestic hot water heating. Such systems show higher performance than conventional water heaters. [9, 13] However, comfort cooling applications of the transcritical CO2 cycle yields lower performance than the typical unitary equipment using conventional refrigerants such as R22, R134a, and R410a [10, 12] because of large expansion losses and higher irreversibility during the gas cooling process. Hence, there have been several researches in such a cycle to optimize its


overall performance by taking the usage of IHX, multi-stage compression, and multi-stage expansion within a system design. So despite the lack of efficiency of the theoretical transcritical cycle, the CO2 supercritical refrigeration cycle may still compete with the vapor compression cycle using other refrigerants. Rather low critical temperature, 31.06°C, and high critical pressure, 73.84 bar, summaries CO2’s main disadvantages. The main drawback of carbon dioxide as a refrigerant is its inherent high working pressure: this pressure is much higher than that of the other natural and synthetic refrigerants. On one hand, this means that for CO2 cycles, newly developed components must be redesigned. Since CO2 offers a much higher volumetric capacity, the problem of the higher working pressure can be overcome by optimal design involving smaller, stronger components. This is the case both in operational situation and standstill condition. However industries have already started facing with success coping with the related problems providing proper safety strategies and components design. Additionally the fact that CO2 has no smell is a disadvantage while high concentration of the CO2 in case of leakage may result a serious problem. So its detection can not be observed without proper detection devices. There is

2 OBJECTIVE

The main objective of this project is to develop, test, and evaluate an energy efficient supermarket system working with CO2 as the refrigerant, emphasize is on using environmentally friendly refrigerants and the choice was to use carbon dioxide and ammonia as the refrigerants in a cascade refrigeration cycle. A laboratory environment allows control over the boundary conditions of system and provides flexibility for modifications. Investigations will focus in first place on overall system evaluation. Detailed analysis of the main components of the system is also an important possibility provided in laboratory environment. Comparison of different cycle modifications will allow optimization of the system for the most efficient solution arrangement, operating conditions and control strategies. [5]

3 CASCADE REFRIGERATION RIG IN IUC

3.1 System Overview

In this investigation the experimental test rig is a cascade system with NH3 at the high stage and CO2 at the low stage, at the medium temperature level CO2 is pumped to provide the required cooling load. Figure 4 is a schematic diagram of the CO2 circuits in the system.


Figure 4 – Schematic diagram of the CO2 circuit in the NH3/CO2 cascade system with CO2 at the medium temperature level

The system solution under investigation has been chosen to be applicable to a medium size supermarket installation in Sweden. The refrigeration system in the supermarkets in Sweden usually operates to satisfy the evaporating temperatures to maintain products at two temperature levels, around +2ºC for cold food and -18ºC for frozen products. Despite the low ambient temperature in Sweden the condensing temperature is usually kept constant around the year at a value of about 40 ºC. The cooling capacities of medium size supermarkets are typically around 50 kW for freezing and 150 kW for cooling at the medium temperature level. This estimate is based on contacts with major installers of supermarket systems in Sweden. Accordingly, the system has been designed to operate between the temperature boundaries mentioned above and to provide a cooling capacity which is scaled down while trying to keep a load ratio of about 3. The low temperature side has a rated capacity of 7.4 kW and the medium temperature side was designed to have a capacity of 20 kW.

3.1.1 System solution

In the choice of the CO2 system the aim was to develop an efficient system with good cooling performance, safe and in accordance with the regulations on the use and release of refrigerants. From environmental point of view, the amount of synthetic refrigerants that can be used to fill the system is limited and the high taxation to prevent leakage makes it expensive to use. For natural refrigerants such as ammonia and propane there is always the safety concern when used in applications where people might be exposed to a leakage accident. CO2 is considered as a relatively safe refrigerant and


classified in group A1. [14] CO2 gas that is used in refrigeration is a by-product of the chemical industry and its use in the refrigeration system can be considered as a delayed step before its unavoidable release to the environment. Therefore, CO2 becomes an interesting solution as a refrigerant from environmental and safety points of view especially in supermarket refrigeration systems where large quantities of refrigerant are required and direct contact, in case of leakage accident, with large number of people might occur. The system that has been chosen is a cascade system with NH3 at the high stage and CO2 at the low stage, at the medium temperature level CO2 is pumped to provide the required cooling load. Figure 5 is a schematic diagram of the CO2 circuits in the system.

Figure 5 – Schematic diagram of NH3/CO2 cascade system with CO2 at the medium temperature level

The usage of the cascade system offers the possibility of utilizing two different refrigerants where each refrigerant is selected to fit the operating range. Using NH3 in the high stage means that it will be easy to deal with a leakage accident as ammonia can only leak into the machine room which should be equipped with proper safety devices. Using CO2 in the low stage, results in reasonable operating pressure levels in the CO2 circuit (28 bars at -8ºC). The favorable pressure drop characteristics of CO2 suits this application where long distribution lines are usually needed. Also this implies that the size of the distribution lines is also smaller than for other refrigerants which it reduces the cost of the piping system. The evaporators at the medium temperature stage are flooded with CO2 which is circulated via a pump; this is expected to produce better performance due to the good heat transfer characteristics of the completely wetted evaporator and therefore the evaporator temperature will be higher than if a direct expansion concept has been used.


3.1.2 System layout

The 20 kW initially designed cooling capacity at the medium temperature level is divided over two display cabinets with 5 kW each and the other 10 kW’s had been supposed to be supplied by an electric heater in what is referred to as a load simulator. The installed electric heater at the medium temperature level managed to provide a maximum of 6.6 kW, which makes the maximum total load at the medium temperature level reduces to 16.6 kW. The electric heater load in simulator can provide three load steps of 2.2, 4.4 and 6.6 kW. On the deep freeze side, the load is divided over two freezers with 2.5 kW each and the electric heater can provide a maximum load of 3 kW. The electric heater load can be provided on three equal load steps in a similar way to the medium temperature load simulator. The maximum cooling capacity of the compressor is 7.4 kW. The freezers are equipped with electronic expansion valves. The compressor is a Copeland scroll type with operating temperatures between -37ºC and -8ºC and a displacement of 4.1 m3/h. The accumulation tank has a capacity to contain 180 L of CO2 and is equipped with an electronic level indicator. It can stand a pressure up to 40 bars which corresponds to an operating temperature of about 6ºC. The system is equipped with a safety release valve that is triggered when the pressure in the system reaches 38 bars. To avoid the opening of the release valve and the loss of significant charge from the system, a bleed valve is installed which opens for periodical release of CO2 at lower pressure than the set value for the release valve, 35 bars, so the pressure in the system will be reduced. If the pressure increase in the system is higher than the rate that the bleed valve can handle, then the release valve will open and release the system’s charge. The CO2 pump that is used is a hermitic one with capacity higher than the highest circulation rate desired; therefore a by-pass is used to reduce the flow rate pumped into the medium temperature circuit. About 1.5 meters head over the pump is respected to prevent cavitations. The ammonia unit uses a Bock reciprocating compressor with displacement of 40.5 m3/h. It can run at 50% reduced capacity by unloading half of its cylinders. Heat is removed from the ammonia evaporator via a thermosyphon loop which required a certain height of the unit. The capacity control of both compressors is achieved by a frequency converter. The cascade condenser is a plate type heat exchanger that is specially selected to handle the pressure difference that will exist between CO2 and ammonia, at -8ºC CO2 will have about 28 bars while ammonia will have a pressure of about 2.7 bars at -12ºC. The medium temperature display cabinets are defrosted using the conventional electric defrost method. The deep freeze cabinets will be


defrosted using hot gas defrost by passing the hot gas through the cabinet, since the condensing temperature of the CO2 at the compressor discharge pressure will be low then heating of the evaporators will be achieved via the sensible heat of the hot gas. Figure 6 is a detailed schematic of the test rig with most of the measuring points indicated. As can be seen in the schematic several by-pass lines and the high number of valves indicate the possible variations in the system which will be used for testing and modifications. The load simulators with the electric heaters can be seen in the diagram in the medium and low temperature circuits. The electric heater provides heat to a brine loop which exchanges the heat with the refrigerant in a plate heat exchanger. The schematic shows that one of the freezers is electrically defrosted while on the rig both freezers are equipped with the hot gas defrost.

Minimum flow nozel

Safety pipe

p

p

M

M

t

t

pt

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Oil drain

p

Highe and low level cutout

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tt p

p

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ttdP

tdP

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Leakage test point

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t

El Defrost

dP

t

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t

t t

dP

p

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dP

t

Figure 6 – Detailed schematic diagram of the NH3/CO2 cascade system test rig

4 MEASUREMENT FACILITIES

The measurement devices consist of below main parts:

• Agilent data acquisition device • Frequency converters • ClimaCheck


• Coriolis flow meter • CO2 compressor power recording • Pulse power meter

Thermocouples and pressure transducers have been connected to various points in the system and through Agilent data acquisition device it is possible to track the system trends on computer. The capacity of ammonia and CO2 compressor varies according to the cooling load. Frequency converters provide a control strategy through a PID controller. It is also possible to record a limited duration trend of each parameters related to the compressors depending on defined scan time interval. In order to record the overall trend of the refrigeration cycle, a set has been applied which is called ClimaCheck. It has been specially designed to record online detailed information of the components during the field measurements or laboratory tests. Performance and detailed information on all key components in the system are calculated, stored and presented in a user friendly style in Microsoft Excel templates. It has been configured accordingly in different stage of the experimental tests and measurements. Its principle to evaluate the system is assuming efficiencies, and measuring the temperatures and pressures around the compressor then it is calculating the mass flow of refrigerant. Based on refrigerant mass flow rate and known measured pressures and temperatures everything else is evaluated. Coriolis flow meter is used to measure the refrigerant flow in medium temperature stage. The Coriolis type is a quite accurate flow meter but the measurement span is from 0 to 1556 g/sec while in our measurements the mass flow rate had a variation between 16 g/sec to 300 g/sec. The accuracy of measurements, in a span less than 5% of total scale, would decrease rapidly. Figure 7 has been taken from mass flow meter manual showing this behavior.

Figure 7– Error in % of actual mass flow rate with 95% confidence (probability)

Looking to the following equation given by manufacturer the zero adjustment plays an important role in flow meter precision:

22 )100()10.0(qm

zE ×+±= (1)


[ ][ ]

[ ]hkgflowMassqmhkgerrorpoZeroZ

ErrorE

/ / int

%

===

But the zero adjustment was unable to perform due to heat gain through the pipes. That made liquid CO2 to evaporate in the pipes and formed the bubbles distorting zero adjustment. To measure the power consumptions in the load simulators and CO2 compressor, the pulse power meters have been connected to the supply electricity in the related component. The analysis and calculations have been based on the CO2 compressor power consumption. It is the same way that ClimaCheck is analyzing a cycle. The pulses counted in a certain time to verify the power consumption in each step of the load simulators. So in this way, each step in medium load simulator verified to be 2.2 kW which resulted in 6.6 kW in total in three steps. The same for freezer load simulator with each step 1 kW and in total three steps added to 3 kW.

5 EXPERIMENTAL RESULT

5.1 Overall system analysis

The system under investigation is a scaled down real installation. The main discussion is weather or not this system is a suitable replacement for traditional technologies. In order to provide answers about the current system solution, it is important to perform the overall analysis of the system where: the capacities are properly measured, the energy balance is verified, and the system’s efficiencies are calculated according to the measurements. This is an important step since some of the measurements are based on the components data and planned experiments depend on the accuracy of measuring cooling loads and capacities of some of the main components.

5.1.1 Load Measurements

The two compressors are used to determine the mass flow rate of the refrigerant which will then be used to calculate the cooling capacities in the corresponding circuits. Initially for both compressors, the compressor manufacturer data were used as guidelines for the calculations which are based on knowing the geometry and the efficiencies of the compressors at certain operation conditions. Measuring the rotational speed of the compressor, and the temperatures and pressures around it gives all the data needed to calculate the mass flow of the refrigerant. Consequently, it will be possible to calculate the energy consumption of the compressor and the cooling capacities of the evaporators/cabinets. Later on, due to the uncertainty with CO2 compressor manufacturer data and its very low performances, CO2 compressor power consumption measured to evaluate


the refrigerant mass flow rate in low temperature and the related cooling capacity. At the return line of the medium temperature level the flow is a two phase, therefore it is not possible to calculate the load at the medium temperature by measuring the mass flow of the refrigerant. By calculating the cooling capacity at the cascade condenser and for the low stage cabinets it will be possible to calculate the total load at the medium temperature level.

5.1.2 Energy Balance Test

The simulators at the medium and low temperature levels provide a fixed known cooling capacity via the electric heaters which can be used to verify the method of calculating the cooling capacity at the medium and low temperature levels using the compressors manufacturers’ data. The medium temperature simulator provides a maximum of 6.6 kW, and the low temperature simulator provides a maximum of 3 kW. The two simulators can be switched to 1/3 and 2/3 of its capacity. The system is run with only the simulators on in addition to the CO2 pump and compressor. The blue line in figure 8 shows the active lines and components in this test.

p

p

M

t

t

t

t

t

Oil drain

p


t

tt p

p

6,6 kW el

ttdP

tdP

t

Leakage test point

t

2.5 kW

t

El Defrost

dP

t

2.5 kW

Hot gas Defrost

t

t t

dP

p

t

dP

t

dP

t

t

t

Figure 8 – Schematic of the active lines and components in the energy balance test

The electric power consumption of the CO2 pump is measured and it varied around the average value of 0.85 kW. The power consumption of the CO2 compressor was measured by an electric meter instead of using the manufacturers’ data, this is due to the fact that the isentropic and volumetric


efficiencies were much less than the provided data. Calculating the mass flow using the compressor data resulted in much higher cooling capacities than the 3 kW provided by the simulator. The compressor was running at a constant rate and the power consumption was measured to be around 1.7 kW. The system is operated around 32ºC for condensing ammonia, -26ºC for freezers, and a medium temperature of about -9ºC, a plot for the temperatures of the boundary conditions of the system during the test period is presented in figure 9.

Evaporating and Condensing Temperature

-30

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-5

0

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pera

ture

(C)

Cond AmmoniadT Cascade CondCond CO2Evap AmmoniaEvap CO2

Figure 9 – Plot of the temperatures at the system boundaries during the energy balance test period [5]

First the system was run with only the medium temperature load (simulator and pump); region 1 shown in figure 10, and then the low temperature circuit is switched on where the low temperature simulator and the compressor power capacities are added to the load at the medium temperature, region 2. The ammonia compressor run at reduced volume of 50% by unloading two cylinders and this is the volume that have been used along the test except in the regions 2.1 and 2.3 where the compressor was switched to full volume. Running the compressor at full volume resulted in occasional stop start operation where the load seemed to be lower than the lowest load to maintain continuous operation of the compressor while switching to half of the compressor effective volume region 2.2 showed that the compressor was running at full speed and was not able to reduce the pressure in the tank to the set value. In order to maintain continuous operation of the compressor the load at the medium stage was reduced to 2/3 of the total capacity with half of the ammonia compressor volume, region 3. Further reduction of 1/3 on the medium temperature was performed in region 4 after which the simulator was switched off. The CO2 pump at the medium temperature level was kept


running in region 5 and then switched off in region 6. In region 7 the low stage was switched off and the system was running with only the CO2 pump on.

Energy Balance

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Q cascade condenserInput capacitydQ(Losses+Comp Ineffec)dQ_average

1

2

2.1 2.2 2.3

3

4

56

7

Figure 10 – Energy balance with fixed value of volumetric efficiency [5]

Running the system at different capacities aims at verifying that the cooling capacity calculated by the ammonia compressor matches different capacities in the system.

5.1.2.1 Cascade condenser cooling capacity

The ammonia mass flow that is passing through the cascade condenser is calculated using the ammonia compressor data, the following equation is used to calculate the refrigerant mass flow of the compressor:

inss Vm ρη ⋅⋅= && (2) Where is the swept volume flow in m3/s, sV& inρ is the density of the refrigerant (kg/m3) at the inlet of the compressor, sη is the volumetric efficiency of the compressor. The compressor has a displacement ( ) of 40.5 m3/hr at rated speed ( ) of 1450. Swept volume flow in m3/s at a given speed can then be calculated using the following relation:

srV&

rn

36001

⋅⋅=r

srs nnVV && (3)

Where is the compressor speed in RPM (1/min). n


The volumetric efficiency of the compressor was extracted from BOCK software by running the software at different operating conditions, for the same operating conditions the mass flow is calculated for an ideal compressor, the ratio of the two values is the volumetric efficiency. An average value for different operating conditions was found to be about 85%. Measuring the compressor speed, the temperature and pressure at the inlet of the compressor provide the required data to calculate the mass flow. The conditions, before and after the cascade condenser, are also measured and therefore the cooling capacity is calculated using the following relation:

dhmQ ⋅= && (4) Where is the enthalpy difference across the heat exchanger. dh In figure 10, the cooling capacity of the cascade condenser is plotted and the difference between this capacity and the provided load is also plotted as dQ which is also presented with average values (dQaverage) over the different operating regions. Excluding region 2 where it was hard to reach the set point and load peaks have been observed at transition time, the difference in load which varies between 1 and 1.8 could be explained as heat losses in the system and deviation in the volumetric efficiency from the estimated value from BOCK’s software. The volumetric efficiency value that was used in the calculations presented in figure 10 is assumed to be constant, which is not the case in practice and it will vary with the pressure ratio. In order to correlate the volumetric efficiency to the operating conditions, a relation suggested by Pierre (1982) for “good” ammonia reciprocating compressor is used to calculate the volumetric efficiency as follows:

⎟⎟⎠

⎞⎜⎜⎝

⎛⋅−⋅=

2

1063.0exp02.1PP

sη (5)

Where 2

1

PP is the pressure ratio.

Using the volumetric efficiency calculated from equation 5 to calculate the mass flow and the cascade condenser capacities yields the results in figure 11. It can be seen from the figure that the difference in cooling capacity is reduced due to the reduced value of the volumetric efficiency and the value varied around 1 kW with smaller deviation around the average value compared to the trend in figure 10.


Energy Balance: calculated volumetric effeciency

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Q Cascade CondenserInput LoaddQdQ_average

Figure 11 – Energy balance with calculated value of volumetric efficiency [5]

If the heat sink into the system is neglected assuming that the difference in the load is due to lower volumetric efficiency than the calculated value then it will be possible to calculate how much should the “actual” volumetric efficiency of the compressor be by using the known load as the input value to calculate the mass flow of refrigerant from equation 4. Consequently, the volumetric efficiency can be calculated using equation 2. The plot in figure 12 shows the volumetric efficiency calculated using the relation in equation 5 and the “actual” volumetric efficiency. As can be seen in the plot the difference between the calculated and actual values is about 10% less for the actual efficiency. Therefore it will be possible to reduce the value calculated in equation 5 by 10% in order to adjust the calculated volumetric efficiency to be closer to the actual value. The third plot in figure 12 is the adjusted value of the volumetric efficiency.


Ammonia compressor volumetric effeciency

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Effe

cien

cy

Calculated

Actual

Adjusted

Figure 12 – Volumetric efficiency of the ammonia compressor [5]

Using the adjusted value for the volumetric efficiency brings the average value for the difference in cooling capacities closer to zero as can be seen in the figure below.

Energy Balance: adjusted volumetric effeciency

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Figure 13 – Energy balance with adjustment to the calculated value of volumetric efficiency [5]

Due to the fact that the used method of calculating the volumetric efficiency is based on estimates and approximations then it is recommended to use another method which is: measure the mass flow of water on the ammonia condenser and measure the temperatures and calculate the condenser capacity. Measuring the compressor power will lead to knowing the cascade condenser cooling capacity.

5.1.2.2 Ammonia compressor capacity and efficiencies

The ammonia compressor is used to estimate the cooling capacity of the cascade compressor and the therefore it is important to evaluate its performance and capacities. Pierre (1982) suggests a relation to estimate the isentropic efficiency ( kη ) for the same compressor relating it to the volumetric efficiency in below equation.

⎟⎟⎠

⎞⎜⎜⎝

⎛+⋅−= 97.169.1exp

2

1

TT

k

s

ηη (6)

The temperature ratio is the absolute temperatures, in Kelvin, of condensation and evaporation corresponding to exist and inlet compressor pressure. Figure 14 shows the calculated and measured values of the isentropic efficiency. Using BOCK software the average isentropic efficiency obtained is about 76% over a range of different operating conditions, which is close to the calculated one, around 78%. The higher value of the measured efficiency may have to do with position of the temperature sensor at the discharge line; it may sense lower temperature value than the actual one. It may also indicate that the volumetric efficiency is higher than the one calculated.


Ammonia Compressor Isentropic Effeciency

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Effe

cien

cy

CalculatedMeasured

Figure 14 – Measured and calculated isentropic efficiencies of the ammonia compressor [5]

Another method that can be used to calculate the cooling capacity is to measure the electric power consumption of the compressor and the enthalpy difference across the compressor is determined by measuring the pressures and temperatures across it. Certain electric motor efficiency and heat losses to the environment should be assumed in order to calculate the shaft power according to the following equation:

elelthermalShaft EE && ⋅⋅= ηη (7) Referring to Climate Check a 7% of thermal losses ( %93=thermalη ) is usually assumed. Usually the shaft power is considered to be the power that is provided to the refrigerant and the mass flow can be calculated according to the equation below, the mass flow is then used to calculate the cooling capacity in equation 4.

compshaft dhmE ⋅= && (7) Figure 15 shows typical efficiency values for well performing electric motor related to the shaft power. The shaft power in the experiment is ranging between 1.8 and 3.3 kW, as seen in figure 12, according to the figure below the electric motor efficiency is estimated to be around 80% at the rated speed.


1,0

0,9

0,8

0,7

0,6

0,50,2 0,5 1 2 5 10 20 50 100 200 kW

η

ηelm

mE&

Figure 15 – Typical values of electric motor efficiency versus the motor shaft power [6]

In the case when the compressor is running at reduced cylinder capacity then there will be additional losses attached to running two non-productive cylinders, in this analysis these losses are included in the electric motor efficiency, therefore it will be lower than the estimated value. The input electrical power is measured and plotted in figure 16 along with the shaft power calculated from the refrigerant side; the adjusted shaft power is the power that is calculated based on adjusted volumetric efficiency presented in figure 12.

Ammonia Compressor Power Consumption

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Shaft powerShaft Power adjustedElectric Power

Figure 16 – Calculated ammonia compressor shaft power and measured electric motor efficiency [5]


Using equation 7 to calculate the efficiency of the electric motor produces the results in the plot of figure 17. The electric motor efficiency value is dependant on the motor speed which is also shown in the figure.

Electric Effeciency vs RPM

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%

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Electirc Effeciency Electirc Effeciency Adjsuted Ammonia RPM

12.1 2.2 2.3 3 4

Figure 17 – Calculated electric motor efficiency for the energy balance test [5]

In regions 1, 2.2, 3 and 4 the compressor is running at reduced capacity with two unloaded cylinders which will add some mechanical losses; these losses are included in the electric motor efficiency presented above. In addition to that in most of the regions the compressor was running at partial speed which reduces the efficiency of the electric motor. In the end part of region 2.3 the compressor was running at full cylinder capacity. Where the operation was close to being stable the efficiency increased due to less mechanical/friction losses, but still the compressor was running at partial speed which reduces the electric motor efficiency. In another test the compressor was running at higher cooling capacity with all the cylinders active. The motor speed was varying due to test conditions and the electric motor efficiency is calculated in a similar manner to the approach above. The plot in the figure below shows that the electric motor efficiency is higher than the values presented in figure 17. This is mainly due to the fact that all the moving cylinders are producing work. It can be seen more clearly in the figure below how the efficiency changes with the speed of the motor. The trend of change is plotted in figure 19 where it indicates that the efficiency tend to increase with motor speed. The results in the figure should not be read as quantitative; the aim of the plot is show the tendency of behavior.


Electric Effeciency vs RPM

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%

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Electirc Effeciency Electirc Effeciency Adjsuted Ammonia RPM

Figure 18 – Calculated electric motor efficiency for ammonia compressor with all cylinders running [5]

Electric Motor Efficiency vs RPM

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800 900 1 000 1 100 1 200 1 300 1 400 1 500RPM (1/min)

Elec

tric

Effe

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cy (%

)

Figure 19 – Calculated electric motor efficiency for ammonia compressor at different motor speeds [5]


5.1.2.3 CO2 compressor capacity and efficiencies

As explained before initially the same method that is used with ammonia unit to calculate the mass flow of refrigerant from the compressor manufacturer data was used for the old CO2 compressor; the results showed a large deviation from the expected values, the resulting cooling capacity was much higher than the provided 3 kW load. This indicates that the actual volumetric efficiency of the compressor is lower than the 90% value used from the manufacturer data. Therefore, the CO2 compressor power that is used in the energy balance was measured by the electric energy meter; the average recorded value was 1.62 kW which can be considered as constant since the load at the low stage was constant and equal to the electric heater capacity of 3 kW. Knowing the capacity and the conditions around the freezer simulator the mass flow of refrigerant can be calculated and then can be used to calculate the power consumption of the CO2 compressor. This way of calculating the compressor power resulted in lower value of about 15% than the measured electric power consumption. This can be seen in the figure below.

CO2 Compressor Power Consumption

-0,20

0,00

0,20

0,40

0,60

0,80

1,00

1,20

1,40

1,60

1,80

13:55:12 14:24:00 14:52:48 15:21:36 15:50:24 16:19:12 16:48:00 17:16:48

kW

CalculatedIsentropic Eff=60%Average Measured

Figure 20 – Calculated old CO2 compressor shaft power and measured electric motor efficiency [5]

The reason for the deviation may be partly due to the difference in the actual discharge temperature and the measured value. The point where the temperature is measured is very close to the compressor exit but due to the very high discharge temperature and the big difference with the room temperature it might be possible that some of the deviation is due to this


difference. In order to estimate how much the actual discharge temperature should be, the measured compressor power is used to calculate the enthalpy at the exit of the compressor and consequently the temperature can be calculated by using the discharge pressure. Figure 21 shows the measured and the expected actual value.

CO2 compressor discharge temperature

50

60

70

80

90

100

110

120

130

140

150

160

170

180

190

200

14:24:00 14:52:48 15:21:36 15:50:24 16:19:12 16:48:00 17:16:48

Tem

pera

ture

C

T_hg_calcT_hg_meas

Figure 21 – Calculated and measured old CO2 compressor discharge temperature [5]

In order to evaluate the isentropic efficiency of the compressor the manufacturer’s data is compared to the measured ones. Figure 22 shows the curve from the manufacturer data where the isentropic efficiency at the running pressure ratio of about 1.7 results in an estimated isentropic efficiency of 60%, assuming that the curve does not drop sharply before the optimum point. Later on, it turned out that the compressor that we have should have a lower curve with about 10% less efficiency. Using this value to estimate the power consumption of the compressor in case of a “good” compressor results in the power consumption presented in figure 20. It was not possible to reach the optimum condition for the pressure ratio due to the fact that reducing the evaporation pressure resulted in the discharge temperature increasing to very high levels.


Isentropic Efficiency

0,40

0,50

0,60

0,70

0,80

0,90

1,00

1,00 2,00 3,00 4,00 5,00 6,00

Pressure Ratio

Figure 22 – Isentropic efficiency of the CO2 compressor at different pressure ratios

Using the calculated mass flow from the simulator side and knowing the compressor geometry and RPM it will be possible to calculate the volumetric efficiency that is presented in figure 23 along with the measured isentropic efficiency which is calculated by measuring the pressures and temperatures around it.


CO2 compressor Isentropic and volumetric effeceincies

10,00

15,00

20,00

25,00

30,00

35,00

40,00

45,00

50,00

13:55:12 14:24:00 14:52:48 15:21:36 15:50:24 16:19:12 16:48:00 17:16:48

%

Isentropic effeceincyVolumetric effeceincy

Figure 23 – Measured Isentropic and volumetric efficiencies of the old CO2 compressor [5]

It is evident that the volumetric and isentropic efficiencies of the compressor are low and this is due to the fact that the compressor have been operating with unfavorable conditions during early runs of the system. It has been notices that the valve pointed out in figure 24 was leaking even when it was firmly closed. When the medium temperature circuit was running with the low stage off liquid CO2 leaked into the oil separator which resulted in bad lubrication for the compressor. This may have created damage in the compressor which was running with normal sound after the valve was replaced and provided the required cooling capacity but at a higher cost of energy consumption due to the low volumetric and isentropic efficiencies.


Minimum flow nozel

Safety pipe

p

p

M

M

t

t

pt

t

t

t

t

t

t

Oil drain

p


p

t

tt p

p

6,6 kW el

ttdP

tdP

t

Leakage test point

t

2.5 kW

t

El Defrost

dP

t

2.5 kW

Hot gas Defrost

t

t t

dP

p

t

dP

t

dP

t

Figure 24 – Schematic diagram of the system shows the leaking line [5]

After changing both the leaking valve and the old compressor, the low temperature circuit was run in evaporating and condensing temperatures which have been recorded along the test according to the below figure.

CO2 Evap and Cond Temperature

-40

-35

-30

-25

-20

-15

-10

-5

0

16:33:36 17:02:24 17:31:12 18:00:00 18:28:48 18:57:36 19:26:24 19:55:12 20:24:00

T Evap CO2T Cond CO220 per. Mov. Avg. (T Cond CO2)20 per. Mov. Avg. (T Evap CO2)

Figure 25 – Evaporating and condensing temperatures for the test in new CO2 compressor


In another measurement, based on formerly mentioned method, once more, cooling capacity was evaluated by measuring the CO2 compressor power. While the power consumption through the Installed electric power meter was measuring by number of pulses during certain time span, the trend in the following figure has been averaged for both cooling capacity and CO2 power consumption along the test.

CO2 Compressor Power Consumption

0,00

2,00

4,00

6,00

8,00

10,00

12,00

17:16:48 17:45:36 18:14:24 18:43:12 19:12:00 19:40:48 20:09:36

kW

CO2 compressor powerCooling capacity20 per. Mov. Avg. (Cooling capacity)

Figure 26 – Power consumption of the new CO2 compressor and the cooling capacity

It should be taken into account that the pulse measurements have been the reason why the values have been jumped along the tests. Looking to the above figure shows that the cooling capacity started from an initial high value of 10 kW and reached below 8 kW after almost 3 hours of test duration. Consequently the CO2 power consumption followed the same trend from slightly below 3 kW to almost 2 kW accordingly. Changing the compressor resulted lower discharge temperature which also shows that the compressor efficiencies have been improved. The test measurements for discharge temperature and improved volumetric and isentropic efficiencies with the new replaced CO2 compressor have been indicated in the following figures.


CO2 compressor discharge temperature

50,00

60,00

70,00

80,00

16:33:36 17:02:24 17:31:12 18:00:00 18:28:48 18:57:36 19:26:24 19:55:12 20:24:00

Tem

pera

ture

C

Figure 27 – The discharge temperature of the new CO2 compressor

CO2 compressor Isentropic and volumetric effeceincies

40,00

50,00

60,00

70,00

80,00

90,00

100,00

110,00

17:16:48 17:45:36 18:14:24 18:43:12 19:12:00 19:40:48 20:09:36

%

2,4

2,5

2,6

2,7

2,8

2,9

3

3,1

3,2

3,3

3,4

Isentropic effeceincyVolumetric effeceincyPressure Ratio10 per. Mov. Avg. (Pressure Ratio)20 per. Mov. Avg. (Isentropic effeceincy)20 per. Mov. Avg. (Volumetric effeceincy)

Figure 28 – Isentropic and volumetric efficiencies of the CO2 compressor with the pressure ratio plotted

Looking to both isentropic and volumetric efficiencies it is clear that it has been considerably improved comparing to the efficiencies in old compressor.


However taking into account the pressure ratio in new case, the compressor has been running on much more favorable condition (figure 22) close to its optimum value for isentropic efficiency. Additionally the variation in pressure ratio from 2.65 to 2.45 in new case verifies its influence on a slight decline to volumetric and isentropic efficiencies. It should be considered that in old compressor, running on higher pressure ratio was impossible because the danger of getting very high compressor discharge temperature which might cause oil burning. That was mainly because of poor old CO2 compressor performances.

5.1.3 System Efficiency

The overall system efficiency is evaluated using the COP which relates the useful cooling load to the work done to provide it. In the earlier test with the old CO2 compressor for the energy balance verification, the cooling capacities are known by running the system on electric heaters in the load simulators as the only load in the system. The power input to the system has been obtained by measuring the electric power consumption of the compressors and pump. This means that all the needed parameters to calculate the COP are available. In another test, after changing the CO2 compressor, the same procedure was applied to evaluate the system COPs and energy balance in addition to running the low and medium cabinets. Based on measured electrical power consumption of CO2 compressor and measured pressures and temperatures over the compressor and its efficiencies, the cooling load in low temperature application is calculated. Then subtracting the pump and low temperature cooling load from cascade condenser capacity gives the rest as the total medium cooling load.

5.1.3.1 Low stage COP

In the early test, the known supplied load to the low stage via the electric heater and the electric power consumption of the old CO2 compressor was measured and the resulting COP is plotted in the figure below and denoted as actual value. The calculated COP in the figure is based on the power consumption that is calculated using the mass flow obtained from the simulator side. Assuming a good compressor that fulfils the manufacturer’s specifications with the estimated 60% isentropic efficiency then the power consumption of the compressor will be much lower with a high COP, as can be seen in the figure.


Low Stage COP

0,00

1,00

2,00

3,00

4,00

5,00

6,00

7,00

8,00

14:09:36 14:38:24 15:07:12 15:36:00 16:04:48 16:33:36 17:02:24

CO

P

Calculated, Eta_v=60%CalculatedActual

Figure 29 – Calculated and measured low stage COP for old CO2 compressor [5]

With the new CO2 compressor that had volumetric and isentropic efficiencies closer to the design values the energy consumption of the compressor was reduced and this improved the low temperature COP comparing to the old one. The below figure is an evidence to the improvement from a low stage COP less than 2 to the new value between 3.5 and 4.


COPs CC 1

1

1.5

2

2.5

3

3.5

4

4.5

5

17:02:24 17:31:12 18:00:00 18:28:48 18:57:36 19:26:24 19:55:12 20:24:00

CO

P

CO2_measuredCO2_calculated10 per. Mov. Avg. (CO2_calculated)

Figure 30 – Calculated and measured low stage COP for new CO2 compressor

5.1.3.2 High stage COP

In the case of the ammonia unit performance evaluation, the capacity of the cascade condenser is considered as the “useful” load which includes the medium, low stage, CO2 compressor, and pump capacities. The COP of the ammonia presented in the figure below is calculated by referring to the directly measured electric power consumption which is referred as “actual”. The “calculated” COP is the one that it is obtained by using the mass flow of refrigerant to calculate the compressor power; the mass flow is based on the calculated volumetric efficiency.


Ammonia Stage COP

0

1

2

3

4

5

6

7

8

9

10

12:28:48 13:04:48 13:40:48 14:16:48 14:52:48 15:28:48 16:04:48

CO

P

Useful cooling loadMeasuredCalculated

Figure 31 – Calculated and measured ammonia unit COP [5]

The measured COP’s of the ammonia unit are lower than expected and this is mainly due to running the system at partial load which is companioned with some mechanical and electric motor losses. At loads close to full capacity the ammonia compressor will have lower electric power consumption, consequently, high stage and total system COP’s are expected to improve.

5.1.3.3 Total COP

The total COP is calculated using two different values for the ammonia compressor power consumption which are discussed in the ammonia unit evaluation above. Also the total COP has been shown for both old and new CO2 compressor in the below figures. However the COP values presented should not be directly compared to each other since the operating conditions and cooling loads are not the same.


Total COP

0

1

2

3

4

5

12:28:48 13:04:48 13:40:48 14:16:48 14:52:48 15:28:48 16:04:48

CO

P

4

5

6

7

8

9

10

CalculatedActualUseful cooling load

Figure 32 – Calculated and measured total COP with the old CO2 compressor [5]

In this test, condensing temperature of ammonia unit was approximately 34ﾟC and evaporating temperature was in average -10ﾟC along the test. Also in average the low temperature cooling had about -30ﾟC as CO2 evaporating temperature and the tank temperature was about -8ﾟC. Corresponding cooling capacities have been also shown that higher loads improve the total COP a little. The test with new CO2 compressor was run with higher capacity while in addition to the load simulators, the freezers and medium cabinets were under operation. The following figure shows the measured cooling capacity in each temperature level all along the test.


4.00

5.00

6.00

7.00

8.00

9.00

10.00

11.00

12.00

13.00

14.00

14:09:36 15:21:36 16:33:36 17:45:36 18:57:36Time

Coo

ling

Cap

acity

(kW

)Q_Medium CoolingQ_Freezer Cooling255 per. Mov. Avg. (Q_Freezer Cooling)255 per. Mov. Avg. (Q_Medium Cooling)

Figure 33 – Medium and Low Stage Cooling Load in the test with new CO2 compressor

Looking to the following figure, condensing temperature of ammonia unit was approximately 34ﾟC and evaporating temperature was in average -10ﾟC along the test. Also in average the low temperature cooling had about -36ﾟC as CO2 evaporating temperature and the tank temperature was about -8ﾟC.

Evap and Cond Temperature

-40-38-36-34-32-30-28-26-24-22-20-18-16-14-12-10-8-6-4-202468

10121416182022242628303234363840

10:48:00 11:16:48 11:45:36 12:14:24 12:43:12 13:12:00 13:40:48 14:09:36 14:38:24 15:07:12 15:36:00

T Evap CO2T Evap AmmoniaT Cond AmmoniaT Cond CO2dT Cascade Cond

Figure 34 – Evaporating and Condensing Temperature for both Low Stage and High Stage in the test with new CO2 Compressor


Improving the efficiencies of the ammonia and CO2 compressors with favorable operating conditions on both stages and with the new CO2 compressor reduced the power consumption and reduced the losses at partial load. This improved the total COP of the system that has been shown in the following figure.

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

5

14:09:36 15:21:36 16:33:36 17:45:36 18:57:36Time

CO

P

CO2AmmoniaTotal

Figure 35 – Measured total COP, High Stage COP, Low Stage COP, with the new CO2 compressor at higher cooling loads

The plotted data are for start up period where the cabinets and products are at room temperature and the cooling capacities are higher than it would be in normal operation. From the experimental values it shows that the COP for the ammonia unit is around 2.7 which is based on the capacity of the cascade condenser and the electric power consumption of the ammonia compressor, while for the low stage CO2 it is around 3.7, and the total COP is about 1.7 and this is based on the total useful cooling load and the total power consumption. The electric motor efficiency is found to be around 80% and with assuming 7% of heat loss from the ammonia compressor then the COP on the refrigerant side (based on shaft power) for the ammonia unit becomes about 3.4 and the total COP is around 2.

5.2 System Variations

In practice, there are many possible variations within the applied CO2 systems, these variations should be investigated in order to modify the system’s design and optimize its cost and improve its performance. In this test rig many variations of the system have been made possible and the design took into consideration the need to modify and adjust some operating parameters.


5.2.1 Gravity Circulation and Pump Circulation Ratio

At the medium temperature side the circulation ratio of the refrigerant can be varied so the pressure drop and heat transfer in the display cabinets and distribution lines can be investigated, low circulation ratios will result in low pressure drop but may influence the heat transfer in the cabinet. Due to change the mass flow rate, using a variable speed pump was impossible due to the fact that it was not possible to find a small pump that can cover the desired flow range. Hence a by-pass line with a regulation valve was used to vary the flow which let more flow go into the by-pass line thus changing the circulation ratio, as. The pump can also be by-passed so the system can be tested for gravity circulation operation. The influence of the circulation ratio on the heat transfer and the pressure drop can be first measured and analyzed over the medium temperature simulator where it will be easier to evaluate the effect on heat transfer by measuring the brine temperatures. The obtained results from the simulator will be used as guidelines to test the proper range of circulation ratios on the display cabinets. The figure below shows the test circuit.

Figure 36 – Schematic diagram of the circulation rate test circuit

The available height in the laboratory is limited to 4 m and the ammonia unit is fixed on top of the CO2 unit due to the thermosyphon loop on the CO2 side of the cascade condenser. The ammonia unit also requires a certain head for its thermosyphon on the cascade condenser side. This means that there is a limited head for the CO2 for gravity circulation, about 1.5 m, which initially considered insufficient to test the system for gravity circulation for full


capacity. However, to have an indication of the possibility of gravity circulation, an investigation for smaller capacities were also performed. Figure 37 shows the temperatures around the heat exchanger during the gravity circulation test. The cooling load was varied from 2.2 kW to 4.4 kW and then to 6.6 kW and as it can be seen in the plot, the mass flow of CO2 is pulsing and temperatures are fluctuating accordingly which indicates that the available height is not sufficient enough to maintain continuous flow of CO2.

Gravity Circulation-Temperatures over the HX in Medium Simulator

-15

-10

-5

0

5

10

15

20

25

9:53:4

3

10:01

:23

10:09

:03

10:16

:46

10:24

:26

10:32

:07

10:39

:51

10:47

:31

10:55

:13

11:02

:53

11:10

:34

11:18

:16

11:25

:56

11:33

:37

11:41

:19

11:48

:59

11:56

:40

12:04

:23

12:12

:03

12:19

:44

12:27

:26

12:35

:06

12:42

:47

12:50

:30

12:58

:10

Time

Tem

pera

ture

ºC

-0.08

-0.06

-0.04

-0.02

0.00

0.02

0.04

0.06

0.08

0.10

0.12

CO

2 M

ass

Flow

(kg/

sec)T_MS1(Brine in)

T_MS2(Brine Out)T_MS_evapT_M2_cCO2_fl (kg/s)

Simulator: (1/3) Simulator: (2/3) Simulator: (3/3)

Figure 37 – Temperatures around the load simulator and the mass flow of CO2 during the gravity circulation test on the load simulator

The pressure drop in the pipes is small and the main pressure drop sources in the CO2 circuit are the filter and the mass flow meter which is not overcome by the limited available head. It is clear that the pressure drops in the filter and mass flow meter depend upon the refrigerant mass flow rate. The average observed value for pressure drops in the filter and mass flow meter at maximum simulator capacity were 0.25 bar and 0.5 bar respectively. They added up to almost 0.75 bar at maximum simulator load. In another test, gravity circulation was tested over one medium cabinet. The figure below shows how the related temperatures and CO2 mass flow rate varied during the test.


Gravity Circulation-Temperatures over the HX in Medium Cabinet 1

-15.00

-10.00

-5.00

0.00

5.00

10.00

15.00

20.00

25.00

13:01

:41

13:05

:31

13:09

:21

13:13

:14

13:17

:04

13:20

:54

13:24

:46

13:28

:36

13:32

:26

13:36

:19

13:40

:09

13:43

:59

13:47

:51

13:51

:41

13:55

:31

13:59

:24

14:03

:14

14:07

:04

14:10

:54

14:14

:46

14:18

:36

14:22

:26

14:26

:19

14:30

:09

14:33

:59

14:37

:51

14:41

:41

14:45

:31

14:49

:24

14:53

:14

14:57

:04

15:00

:56

15:04

:46

15:08

:36

Time

Tem

pera

ture

ºC

-0.06

-0.04

-0.02

0.00

0.02

0.04

0.06

0.08

0.10

CO

2 M

ass

Flow

(kg/

sec)

T_MC1_air_inT_MC1_air_outT_MC1_evap_inT_MC1_evap_outCO2_fl (kg/s)

CO2 mass flow is varying to the great extent !

Back flow of CO2 measured !

Figure 38 – Medium Cabinet Temperatures and the mass flow of CO2 during the gravity circulation test on the medium cabinet

Looking to the trend for supply temperature in the graph shows that passing more than 2 hours could lower the supply air temperature over the product slightly less than 5ﾟC. Of course this is not sufficient for the product temperature where they should be kept at 2ﾟC. The same trend with CO2 mass flow rate observed comparing to the case with simulator while the value was varied between 0.01 kg/sec to 0.08 kg/sec. The medium temperature simulator is used to test for different circulation rates. The fixed load provided in the simulator makes it possible to test of a wide range of circulation ratios. Carioles mass flow meter is used to measure the mass flow of refrigerant, the flow rate at low circulation ratios of about 2 is very small compared to the measuring range of the mass flow meter. Due to the continuous generation of bubbles in the CO2 line it was not possible to perform the zero adjustment reading on the flow meter. For full capacity of the simulator the mass flow that is required to result in a circulation rate of 1 is 0.026 kg/s; less than 5% of the range of the mass flow meter where the estimated error is expected to be high. The valve located before the medium temperature simulator is used to control the mass flowing through the simulator. In order to reach the point of circulation 1 the valve was almost closed so that superheating is achieved in the simulator, it can be seen in the sight glass that there is no two phase flow in the circuit. Then the valve has been opened on small steps until the superheat was eliminated and the sight glass shows no liquid. It was noticed that slight opening of the valve after this point was reached resulted in liquid drops visual in the sight glass. The value of the mass flow reading at that


point was used as the reference value to calculate the circulation ratio. The mass flow showed a reading of about 0.022 kg/s. Circulation rate have been changed between one and slightly over 14, the reason that this high value was reached was that the valve was opened gradually to identify at which point the heat transfer starts to change, the valve was fully open at the maximum circulation rate. Figure 39 shows the pressure drop across the simulator at different steps of circulation rate. As can be seen from the figure, the waiting time for each step is rather short, around 20 minutes, especially for the steps with low circulation rates. In order to confirm the accuracy in the data generated in this test another test have been run where the run time was longer, at least 1 hour, for less steps than the first test, figure 40.

Pressure drop and CR

0,20

0,40

0,60

0,80

1,00

1,20

1,40

1,60

8:48:29 9:24:29 10:00:29 10:36:29 11:12:29 11:48:29 12:24:29 13:00:29 13:36:29 14:12:29 14:48:29 15:24:29 16:00:29

dP (b

ar)

1

2

3

4

5

6

7

8

9

10

11

12

13

14

15

16

CR

dP

CR

Figure 39 – Simulator’s pressure drop and circulation ratio for short waiting time test [5]


Pressure drop

1

3

5

7

9

11

13

15

10:19:1210:55:1211:31:1212:07:1212:43:1213:19:1213:55:1214:31:1215:07:1215:43:1216:19:1216:55:12

dP (b

ar)

0,00

0,20

0,40

0,60

0,80

1,00

1,20

1,40

CR

CRdP

Figure 40 – Simulator’s pressure drop and circulation ratio for long waiting time test [5]

As it can be seen from the figures the pressure drop across the heat exchanger increases with increasing the mass flow of refrigerant. Figure 41 relates the pressure drop as a function of CR, the plots in the diagram shows agreement between short and long run time tests.


Pressure drop vs CR

0,00

0,20

0,40

0,60

0,80

1,00

1,20

1,40

1,60

0,00 1,00 2,00 3,00 4,00 5,00 6,00 7,00 8,00 9,00 10,00 11,00 12,00 13,00

CR

dP (b

ar)

Short Test PeriodLong test Period

Figure 41 – Simulator’s pressure drop at different circulation ratio for short and long waiting time [5]

Running at high circulation rate and having a high pressure drop could be justified if the heat transfer in the evaporator would improve with high circulation rate where an optimum circulation rate may exist. Measuring the brine temperatures around the simulator it was possible to evaluate the influence of the circulation rate on the heat transfer in the heat exchanger. Figure 42 shows the brine temperatures around the heat exchanger, the evaporation temperature of CO2 and the logarithmic mean temperature difference for the test with short waiting time test.


-11

-9

-7

-5

-3

-1

1

3

5

7:12:00 8:24:00 9:36:00 10:48:00 12:00:00 13:12:00 14:24:00 15:36:00 16:48:00

Tem

pera

ture

(C)

LMTDT Brine InT Brine OutT CO2

Figure 42 – Temperatures around simulator [5]

As can be seen from the figure, within the tested range the circulation ratio has almost no influence on the heat transfer in the heat exchanger; LMTD had almost constant value of about 3.6ºC along the test. Plot of the LMTD versus the circulation ratio for the two tests is presented in figure 43 where the influence of the circulation ratio can be seen in a clearer way, the influence is very small where the LMTD caries within a range of 0.2ºC which is insignificant.


LMTD vs CR

3,3

3,4

3,5

3,6

3,7

3,8

3,9

4

0,00 1,00 2,00 3,00 4,00 5,00 6,00 7,00 8,00 9,00 10,00 11,00 12,00 13,00

CR

LMTD

C

Short Test PeriodLong Test Period

Figure 43 – Simulator’s LMTD at different CR for long and short waiting time tests [5]

As it can be seen from the results above increasing the mass flow of refrigerant had an insignificant improvement on heat transfer. Increasing the circulation ratio resulted in an increase in the pressure drop across the heat exchanger without improvement on the heat transfer; this indicates that the circulation ratio should be chosen as low as possible to ensure complete evaporation at the highest load expected. There was no optimum operating circulation ratio found similar to the case of conventional brines in secondary systems, this gives more flexibility in choosing the operating conditions for CO2 secondary system, especially that its pressure drop is much lower than it is for brines. The same configuration as what performed on simulator tested on one medium cabinet. The following figure shows the relation of LMTD and pressure drop to the variation of circulation ratio.


Pressure Drop and LMTD in Medium Cabinet in Different Circulation Rate

-0.30-0.100.100.300.500.700.901.101.301.501.701.902.102.302.502.702.903.103.303.503.703.904.104.304.504.704.90

0.96

0.98

1.01

1.01

1.00

0.97

0.99

0.98

0.98

2.01

2.01

1.99

1.98

2.01

1.99

2.53

2.53

2.53

2.52

3.01

3.00

3.03

3.00

3.04

3.00

3.04

2.04

1.48

1.49

1.51

1.00

0.99

1.01

1.02

0.98

(CR)Circulation Ratio

Pres

sure

Dro

p (b

ar),

LMTD

(ﾟC

)

PD_MC_1LMTD

Figure 44 – Medium cabinet’s pressure drop and LMTD versus different circulation ratio for gravity circulation test

As it can be seen in the figure the pressure drop in the cabinet was not able to measure since it was extremely low in medium cabinet evaporator. That is due to the fact that the cabinet evaporator has not been designed to appropriately work with CO2 as the refrigerant. It means that the pipes have excessive large diameter and haven’t been designed to the optimum performance. Additionally the pressure transducers failed to measure such too small pressure difference around the cabinet however form the trend influence of increasing circulation ratio on pressure drop can be seen. The same evidence exist the same as with the gravity circulation in simulator since the main pressure drops are in mass flow meter and the filter depending on the refrigerant mass flow rate. Changing circulation rate has also insignificant influence improving the heat transfer in medium cabinet. Looking to LMTD trend in above figure shows that increasing circulation rate has just decreased the LMTD from 4.5ºC to 3.5ºC. No optimum value observed all along the test. Also measuring the air temperatures around the medium cabinet was performed. Following figure shows the air temperatures around the medium cabinet heat exchanger, the evaporation temperature of CO2 and the logarithmic mean temperature difference for the test with variation of circulation ratio.


Medium Cabinet with Different CRs

-9.00

-7.00

-5.00

-3.00

-1.00

1.00

3.00

5.00

7.00

9.00

11.00

13.00

13:17

:00

13:27

:50

13:38

:41

13:49

:31

14:00

:23

14:11

:14

14:22

:04

14:32

:56

14:43

:47

14:54

:39

15:05

:30

15:16

:22

15:27

:13

15:38

:06

15:48

:57

15:59

:50

16:10

:40

16:21

:33

16:32

:24

16:43

:14

16:54

:07

17:04

:58

17:15

:50

17:26

:42

17:37

:35

17:48

:26

17:59

:19

18:10

:11

18:21

:04

Time

Cel

cius

T_air_in (Average)T_air_out (Average)LMTDT_evap_inT_MS4_evapCR(Circulation Ratio)

Figure 45 – Temperatures around the medium cabinet

5.2.2 Cascade Condenser

On the cascade condenser totally four main arrangements were tested both in medium temperature circuit and concerning the way CO2 condenses. In each circuit two arrangements were examined as what we called it indirect and direct arrangements. In each test in low and medium circuit, the system was running with almost the same capacity at the cascade condenser. During the earliest tests, the system has been all the time running on the thermosyphon arrangement described in figure 50. Due to the resulted high discharge temperature of CO2 the two other arrangements have been tested after changing the CO2 compressor. The first arrangement is the one in figure 46, where the two phase CO2 return line from the medium temperature circuit ends in the accumulation tank.


Figure 46 – Schematic diagram for indirect or thermosyphon arrangement: (medium return to the tank)

Along the test for medium temperature circuit, the low temperature circuit was abandoned. The system configuration is also called thermosyphon arrangement since the saturated vapor CO2 condenses in the cascade condenser and return back to the tank through the indicated loop in above figure. Figure 47 shows the temperature values across the cascade condenser; the difference between the “hot” CO2 and the “cold” ammonia is almost at 2°C all along the test. It should be noticed that the load was only from the medium temperature side by running two medium cabinets. Measured cascade condenser capacity along the test had an inclination from 12.2 kW to 10.5 kW.

Cascade C ondenser Tem peratures

-12.00

-10.00

-8.00

-6.00

-4.00

-2.00

0.00

2.00

4.00

12:00:00 12:28:48 12:57:36 13:26:24 13:55:12 14:24:00 14:52:48 15:21:36 15:50:24 16:19:12 16:48:00

Tim e

Tem

pera

ture

(C)

dTC O 2 C ondensingAm m onia evaporating

Figure 47 – Temperatures across cascade condenser for indirect or thermosyphon arrangement in medium circuit


The other arrangement in medium temperature circuit is where two phase CO2 return line from the cabinets is directly passing through the cascade condenser. The system is so called the direct arrangement while the CO2 condenses directly in the cascade condenser. Following figure shows the basic arrangement.

Cascade condenser

Cold Food

NH3 unit

CO2

T_evap=-8C

Figure 48 – Schematic diagram for direct or forced condensation arrangement: (Medium return to the cascade condenser)

Following figure shows the temperature values across the cascade condenser; the temperature difference between the condensing CO2 and the evaporating ammonia is again almost constant at 2°C during the test. The same condition applied to the test same as the indirect arrangement. The medium temperature side by was run on two medium cabinets. The measured cascade condenser capacity along the test had an inclination from 12 kW to 11 kW which is almost the same condition as the indirect arrangement.

Cascade Condenser Temperatures

-12

-10

-8

-6

-4

-2

0

2

4

14:38:24 15:07:12 15:36:00 16:04:48 16:33:36 17:02:24 17:31:12 18:00:00 18:28:48 18:57:36

Time

Tem

pera

ture

(C)

Evaporation AmmoniaCondensation CO2dT


Figure 49 – Temperatures across cascade condenser for direct or forced condensation arrangement in medium circuit

Further arrangements were subsequently performed to compare variations on low temperature circuit. In the first low temperature circuit arrangement, CO2 is condensing in a thermosyphon loop, Figure 50, where the two phase CO2 return line ends in the accumulation tank and the hot gas return from the low stage passes through the liquid of the tank so the hot gas will be de-superheated by boiling off some of the tank’s liquid.

Figure 50 – Schematic diagram for indirect or thermosyphon arrangement: (discharge hot gas to the tank)

Figure 51 indicates the temperature values across the cascade condenser; the temperature difference between CO2 and ammonia while they are respectively condensing and evaporating is approximately constant around 3°C during the test. It should be considered that the load was measured in the cascade condenser as a slight decrease from 20 kW to 19 kW.


Thermosyhon Evap and Cond Temperature

-12

-11

-10

-9

-8

-7

-6

-5

-4

-3

-2

-1

0

1

2

3

4

16:33:36 17:02:24 17:31:12 18:00:00 18:28:48 18:57:36 19:26:24 19:55:12 20:24:00

T Evap Ammonia

T Cond CO2

dT Cascade Cond

100 per. Mov. Avg. (dT Cascade Cond)

100 per. Mov. Avg. (T Cond CO2)

100 per. Mov. Avg. (T Evap Ammonia)

Figure 51 – Temperatures across cascade condenser for indirect or thermosyphon arrangement in low temperature circuit

The final tested arrangement, figure 52, is where the hot gas return line from the low temperature level is passing directly through the cascade condenser after mixing with the saturated vapor from the tank.

Cascade condenser

Cold Food

Frozen Food

NH3 unit

CO2CO2

T_evap=-37C

T_evap=-8C

Figure 52 – Schematic diagram for direct or forced condensation arrangement: (Discharge hot gas mixing with the sat. vap. from the tank then to the cascade condenser)

The arrangement in figure 52 has been tested and the results are similar to the ones plotted in figure 51. Figure 53 shows the temperatures across the cascade condenser in new case; the temperature difference between CO2 and ammonia while they are respectively condensing and evaporating is again roughly constant in the order of 3°C during the test. In this test, the load


was measured in the cascade condenser as a slight decrease from 19 kW to 17.5 kW.

CC 2 Evap and Cond Temperature

-12

-11

-10

-9

-8

-7

-6

-5

-4

-3

-2

-1

0

1

2

3

4

10:48:00 11:16:48 11:45:36 12:14:24 12:43:12 13:12:00 13:40:48 14:09:36 14:38:24 15:07:12 15:36:00

T Evap AmmoniaT Cond CO2dT Cascade Cond100 per. Mov. Avg. (dT Cascade Cond)100 per. Mov. Avg. (T Cond CO2)100 per. Mov. Avg. (T Evap Ammonia)

Figure 53 – Temperatures across cascade condenser for direct or forced condensation arrangement in low temperature circuit

Initially it was expected to have better system stability while the thermosyphon arrangement was running. Later on based on the performed tests and looking to the above figures, the system had apparently the stable operation in all variations. In all variations, the temperature drop that is measured and observed while operating the cascade condenser indicates a good heat transfer that yields a low temperature difference across the cascade condenser. This is mainly due to the favorable conditions for exchanging heat on the sides of the heat exchanger. CO2 enters the cascade condenser as saturated vapor in thermosyphon arrangement and ammonia boils off along the heat exchanger from saturated conditions at the inlet. Further evaluation for direct arrangement or forced condensation in cascade condenser showed almost the same performance in heat transfer. That is mainly due to the high heat transfer coefficient in ammonia side while it is boiling. Also higher mass flow rate in case of forced condensation has been a reason to an equal heat transfer performance with the thermosyphon arrangement.

5.2.3 Flashing Gas in Liquid Lines

Long supply lines to the freezing cabinets are usually the case in this application which means that there will be a pressure drop and some of the refrigerant might flash before the expansion valve resulting in unstable operation. An internal heat exchanger is usually used to tackle this problem by adding some sub-cooling before the distribution lines, figure 56. Another


possibility is to add some head to the liquid and this is achieved by connecting the liquid supply to the expansion valves after the medium temperature circulation pump, Figure 57, in this case the additional pumping power accompanying this solution will be very small since the CO2 pump is larger than needed.

Figure 54 – Basic schematic of internal heat exchanger solution

Figure 55 – Basic schematic of the solution where head is added to the liquid before expansion valve

So far the system has been running without pump head or internal heat exchanger. The pressure drop and heat leaking in the liquid supply line proved to be small and the available head in the tank was enough to overcome it. Both arrangements in the figure above have been tested and they work properly, when running with internal heat exchanger the discharge gas temperature was noticed to be very high. As it will be explained in the next section, the minimum superheat value had to be limited on a quite high


value in favor of proper control. This made us to omit internal heat exchanger during the tests.

5.2.4 Freezing Cabinets Control

The two freezers in the system are identical and equipped with electronic expansion valves. The freezers were delivered with a control system that measures the evaporation temperature on the surface of the evaporator tube. In this case the thermocouple is placed at the first bend after the expansion valve. Figure 58 is a schematic diagram of the freezer evaporator where the positions of the measuring points are indicated.

Figure 56 – Basic schematic of the freezer’s evaporator with the measuring points indicated

The freezers showed occasional unstable behavior, especially at start up, while running the system with this control method, this was more likely to happen at low superheat set point. While running the system at the lowest set point of 3°C the air temperature in the freezers started to increase up to a certain value and then decrease again and the behavior is repeated in a cyclic manner. This is plotted in figure 59 where it can be seen that the air temperature difference is decreasing which indicates loss in cooling capacity.


-40

-30

-20

-10

0

10

20

13:48:00 14:16:48 14:45:36 15:14:24 15:43:12 16:12:00

Tem

pera

ture

(C)

Superheat Air in Air out Freezer surface-controller Freeze

Superheat=3°

Figure 57 – Temperatures around the freezer cabinet with temperature based controller. Superheat set value is 3ºC. [5]

The controller starts controlling the system for the highest superheat set value, usually set between 10-15°C, and then the valve opening is adjusted targeting the lowest superheat set value, where the system is usually operating at in a stable condition. When the system reaches the lowest superheat set point this behavior is triggered. This can be seen in the superheat value in the figure above where the controller is trying to control the system to operate at the superheat value of 3°C. During the unstable operation there was not enough refrigerant flowing in the evaporator which led to the refrigerant being superheated at the point where the evaporation temperature should be measured. The evaporating temperature measured by the pressure transducer is plotted in the diagram which indicates the difference between the measurement of the real evaporating temperature from the pressure reading and the one that is measured on the surface of the heat exchanger tube. The superheat set point of 3°C is very low compared to practice but the same behavior was repeated with higher superheat set values but with fewer oscillations after the start up period. A sample of the tests is shown in figure 60 where the superheat set value of 7°C was used. The unstable behavior occurred only once at start up of the test.


-40

-30

-20

-10

0

10

11:24:00 11:52:48 12:21:36 12:50:24 13:19:12 13:48:00 14:16:48 14:45:36 15:14:24

Tem

pera

ture

(C)

T SH SurfaceT Air InT Air outT Freeze SurfaceT Freeze Sat Press

Superheat=7

Figure 58 – Temperatures around the freezer cabinet with temperature based controller. Superheat set value is 7ºC. [5]

Another method of measuring the evaporation temperature is by measuring the evaporation pressure and this will be the input signal to the expansion valve controller. A schematic diagram of the pressure based controller method is presented in figure 61. In this case the actual evaporation temperature is measured instead of the surface temperature which is about 2°C higher than the real value; this can be seen in the plots in the following figures below.

Figure 59 – Basic schematic of the freezer’s evaporator with the measuring points indicated


The controller was changed on one display cabinet while the other was kept on the surface temperature measurement method. The freezers were run at the same conditions and at the same set point. The influence of the superheat value is tested, a value of 8°C was used and the two cabinets were running with good stability as can be seen in figure 62 which shows the air and brine temperatures for the cabinets and the simulator. The simulator used the pressure based control all the time.

Freezers and Simulator Temperature

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-10,00

0,00

10,00

20,00

0,00 100,00 200,00 300,00 400,00 500,00 600,00 700,00

Tem

pera

ture

(C)

Air in Press ControlAir out Press ControlAir in Temp ControlAir out Temp ControlBrine in Press ControlBrine out Press Control

Superheat=8-10°C

Time span 09:10:34-11:12:43

Figure 60 – Temperatures around the freezers and simulator. Superheat set value is 8ºC. [5]

Figures 62 and 63 shows evaporating temperatures, superheat values and the superheat temperature measured at the exit of the evaporator for the freezers and the simulator. The variation in the superheat temperature in the freezer with surface temperature control, in figure 62, may be due to the fact that the surface temperature controller senses lower superheat than the pressure one and starts regulate at earlier time than the pressure based controller. As can be seen in figure 62, the measured superheat temperature at the exit of the evaporator in case of the simulator is fluctuating at start up which may be due to the small size of the heat exchanger where the response to the expansion valve opening or closing is faster and more prominent than the case of the long freezer heat exchanger.


Freezers Controlers Paramters

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0,00

10,00

20,00

30,00

40,00

0,00 100,00 200,00 300,00 400,00 500,00 600,00 700,00

Tem

pera

ture

(C)

SuperHeat Press ControlSuperHeat Temp ControlEvap exit Press ControlEvap exit Temp ControlFreezing Temp ControlFrezer Press Control

Superheat=8-10°C

Time span 09:10:34-11:12:43

Figure 61 – Temperature values input to the controller in the freezers. Superheat set value is 8ºC. [5]

Simulator Controlers Paramters

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-10,00

0,00

10,00

20,00

30,00

40,00

0,00 100,00 200,00 300,00 400,00 500,00 600,00 700,00

Tem

pera

ture

(C)

Super HeatEvap exitFrezering

Superheat=8-10°C

Time span 09:10:34-11:12:43

Figure 62 – Temperature values input to the controller in the simulator. Superheat set value is 8ºC. [5]

When the superheat temperature is reduced to 5°C in an attempt to trigger the unstable behavior both freezers started to show cyclic increase and decrease


in the air temperature, figures 65 and 66 shows the air and superheat temperatures on both cabinets. In case of surface temperature control the oscillations had higher peaks and the air exit temperature was close to 0°C, while in the other case the value did not go over -12°C. And the superheat in the second case is oscillating between 2 and 10°C while in the first case the superheat value goes below zero and then increase up to 20°C. The superheat high set value for the controller is 10°C in both cases.

Freezers Temperature and Superheat

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Tem

pera

ture

(C) Superheat

Air inAir out

Superheat=5-10°C

Figure 63 – Temperatures around the freezer with pressure based controller. Superheat set value is 5ºC. [5]


Freezers Temperature and Superheat

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Tem

pera

ture

(C) Supeheat

Air inAir out

Superheat=5-10°C

Figure 64 – Temperatures around the freezer with temperature based controller. Superheat set value is 5ºC. [5]

The parameters that the controllers read in both cases are presented in figures 67 and 68, as can be seen in figure 67 the evaporating temperature read is increasing due to refrigerant superheat and the controller reads the wrong evaporation temperature and fails to control properly.

Freezers Controllers Parameters

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-10

-5

0

5

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11:38:24 12:07:12 12:36:00 13:04:48 13:33:36 14:02:24 14:31:12 15:00:00

Tem

pera

ture

(C) Freezer Press Control

Super HeatFreezer Temp ControlEvap Exit

Superheat=5-10°C


Figure 65 – Temperature values input to the pressure based controller in the freezer. Superheat set value is 5ºC. [5]

Freezers Controllers Parameters

-30

-25

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-15

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-5

0

5

10

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11:38:24 12:07:12 12:36:00 13:04:48 13:33:36 14:02:24 14:31:12 15:00:00

Tem

pera

ture

(C) Freezer Press Control

Super HeatFreezer Temp ControlEvap Exit

Superheat=5-10°C

Figure 66 – Temperature values input to the temperature based controller in the freezer. Superheat set value is 5ºC. [5]

In the case of pressure based controller it measures a higher superheat value than the actual one, which can be seen in the figures above as a difference between surface measured evaporation temperature and the pressure related one. This is true assuming that the same temperature difference exists at the superheat measuring point. This gives the pressure based controller higher margin of stability, on the other hand the temperature based controller may be closer to measuring the real superheat value but in the cases discussed above at some conditions it measures wrong evaporation temperature. Also the response to changes or fluctuations in evaporation temperature would be faster in case of the pressure based controller. In the case of the simulator the controller managed to regulate the superheat value adequately at the 5°C set value, this can be seen in figure 69. This may be due to the fast response at the evaporator exit to any changes to the degree of opening of the expansion valve which consequently readjusts itself accordingly.


Freezer Simulator Brine Temperature and Superheat

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11:45:36 12:14:24 12:43:12 13:12:00 13:40:48 14:09:36 14:38:24 15:07:12

Tem

pera

ture

(C)

SuperheatBrine inBrine out

Superheat=5-10°C

Figure 67 – Temperatures around the simulator with superheat set value of 5ºC. [5]

The measurements and experiences in running the display cabinets with the two controllers indicate a more favorable conditions in case of the pressure based controller. The controller successfully kept the superheat value within the set range, it had lower stability at low superheat set values but still not as unstable as the temperature based controller. Measuring the evaporation temperature on the surface of the tube of the evaporator proved to be the source of instability in the controller; the point measured superheated vapor when the mass flow was small in the heat exchanger.

6 CONCLUSION AND DISCUSTION

The chosen NH3/CO2 system for supermarket refrigeration is being built in a laboratory environment which can replicate the conditions in a real installation. Tests have been run for overall evaluation of the system and for the main components. The measured value for COP of the system gives in average about 5% off the computer simulation model. The deviation may be due to the fact that the theoretical model does not consider the heat losses into the system, pressure drops and the variations in superheat value at the freezers. Also there might be inaccuracies in calculating the compressors’ efficiencies and in the manufacturers’ data. In gravity circulation it is possible to omit the expensive special hermetic CO2 pump. In the existing test rig, the CO2 pump capacity was exceptionally higher than required. This was due to the fact that the pump within our required low capacity was unavailable. Considering the limited available


height in the test rig, the gravity circulation failed to meet the cooling load. This might be the case in most of the real installations too. Hence designing supermarket refrigeration based on gravity circulation is absolutely site dependent and it needs careful design consideration. Since the CO2 pressure drop in the system is quite low, the pump just needs to overcome a quite low pressure drops. So while the CO2 refrigeration market is expected to grow in future, the manufacturers would certainly provide broad ranges of the capacity. It would make the designers’ choice extensive as they would be able to cope with their required range of the pump capacity. It is also expected that designing gravity circulated system and dealing with the site specific considerations and its troubles wouldn’t worth to the pump circulated system. The circulation ratio in the simulator and medium temperature cabinet has been changed and its influence on heat transfer and pressure drop has been evaluated. Increasing the mass flow of refrigerant to have circulation ratios higher than 1, had insignificant improvement on heat transfer. The pressure drop across the heat exchanger and medium cabinet increased and no optimum operating point have been found. Therefore, the circulation ratio should be chosen as low as possible to ensure complete evaporation at the highest expected cooling demand. No difference in heat exchanging performances observed between different arrangements for the cascade condenser. The measured temperature drop while operating the cascade condenser in all variations indicates superior heat transfer conditions. This is mainly due to the favorable conditions for exchanging heat on both sides of the heat exchanger. In thermosyphon arrangement CO2 enters the cascade condenser as saturated vapor and ammonia boils off along the heat exchanger from saturated conditions at the inlet. In case of direct arrangements or forced condensation for cascade condenser, higher mass flow may resulted an equal heat transfer performance with the thermosyphon arrangement. Surprisingly the tests showed that the system is quite stable in all variations. The measurements and experiences in running the display cabinets with the temperature and pressure based controllers indicated more favorable conditions in case of the pressure based controller. For minimum superheat set values higher than 6°C, the controller successfully kept the superheat within the set range, it had lower stability at lower minimum superheat set values but still not as unstable as the temperature based controller. Also while keeping the stable operation of the system in low superheat value was hard to achieve, the suction line temperature to CO2 compressor had not the opportunity to lower the inlet temperature before the expansion valve in IHX. In other words, the use of IHX was not applicable to improve the low stage COP. The possibility of improving the low stage COP through an IHX would be practicable in case of fine superheat control strategy.


Finally, CO2 technology is attractive as it is environmentally benign and locally safe. It just needs a breakthrough enabling mass production of the necessary specific components like appropriate superheat controllers through a suitable strategy in CO2 application. It is also the case for hermetic CO2 pump as it should let the designers to have broad range to select. Also they should be cost-competitive compared with conventional refrigeration, air-conditioning and heat-pump technology with either inherent global environmental or local safety problems.

7 REFERENCES

[1] Sawalha, S., Soleimani, A., Rogstam, J., Experimental and Theoretical Evaluation of NH3/CO2 Cascade System for Supermarket Refrigeration in a Laboratory Environment, 7th IIR Gustav Lorentzen Conference on Natural Working Fluids, Trondheim, Norway, (May 2006). [2] Wikipedia contributors, "Refrigerator," Wikipedia, The Free Encyclopedia, http://en.wikipedia.org/w/index.php?title=Refrigerator&oldid=46395378 (accessed April 1, 2006). [3] Wikipedia contributors, "Supermarket," Wikipedia, The Free Encyclopedia, http://en.wikipedia.org/w/index.php?title=Supermarket&oldid=45602508 (accessed April 1, 2006). [4] Pearson, A., Carbon dioxide—new uses for an old refrigerant, Int. J Refrigeration 28, 1140–1148, (2005). [5] Sawalha, S., Soleimani, A., Rogstam, J., CO2 in Supermarket Refrigeration, 1st phase project report, IUC, (Dec. 2005). [6] Sawalha, S., Using CO2 in Supermarket Refrigeration, ASHRAE Journal, Vol. 47, No. 8, 26–30, (Aug. 2005). [7] Arias, J., Energy Usage in Supermarkets Modelling and Field Measurements, Doctoral thesis, Department of Energy Technology, Royal Institute of Technology, Stockholm, Sweden, (2005). [8] Granryd, E., Ekroth, I., Ludqvist, P., Melinder, Å., Palm, B., and Rohlin, P., Refrigeration Engineering, Department of Energy Technology, KTH, Stockholm, Sweden, (2003). [9] Neksa, P., CO2-heat pump systems, Int J Refrigeration 25, 421-427, (2002). [10] Hwang, Y., Huff, H., Preissner, R., Radermacher, R., CO2 transcritical cycles for high temperature application, proceedings of 2001 ASME international mechanical engineering congress in New York, (2001) [IMECE2001/AES-23630].


[11] Zhao, Y., M. Molki, et al., Flow Boiling of CO2 in Micro channels. ASHRAE Transactions 106(1), 437-445, (2000). [12] Hwang, Y., Radermacher, R., Experimental investigation of the CO2 refrigeration cycle, ASHRAE Trans 105(1), 1219-1227, (1999). [13] Neksa, P., Rekstad, H., Zakeri, R., Shiefloe, P., CO2-heat pump water heater: characteristics, system design and experimental results, Int J Refrigeration 21(3), 172-179, (1998). [14] American Society of Heating, Refrigerating, and Air-Conditioning Engineers, ASHRAE handbook of refrigeration (Fundamental), ASHRAE, Atlanta, USA, (1997). [15] Lorentzen, G., Revival of carbon dioxide as a refrigerant, Int. J. Refrigeration Volume 17 Number 5, 292–301, (1994). [16] Lorentzen, G., Trans-critical vapour compression cycle device Patent No. WO/07683, (1990). [17] ASHRAE Standard 15-2004 -- Safety Standard for Refrigeration Systems [18] EN378: 2000: Refrigerating systems and heat pumps – safety and environmental requirements

EXPERIMENTAL INVESTIGATIONS OF NH3/CO2 CASCADE SYSTEM … · variations of the system solution are...

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Transcript of EXPERIMENTAL INVESTIGATIONS OF NH3/CO2 CASCADE SYSTEM … · variations of the system solution are...